Patent Publication Number: US-10329970-B2

Title: Custom VVA rocker arms for left hand and right hand orientations

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     This application is a continuation of U.S. patent application Ser. No. 14/876,026, filed Oct. 6, 2015, which is hereby incorporated by reference in its entirety. 
     U.S. patent application Ser. No. 14/876,026 is a continuation of U.S. patent application Ser. No. 14/188,339, filed Feb. 24, 2014, which is hereby incorporated by reference in its entirety. 
     U.S. patent appl. Ser. No. 14/188,339 claims the benefit of U.S. Provisional Patent Application 61/768,214 filed Feb. 22, 2013, entitled “Custom VVA Rocker Arms for Left Hand and Right Hand Orientations.” 
     All of the foregoing patents and patent applications are incorporated herein in the entirety for all purposes. 
    
    
     FIELD 
     This application is related to rocker arm designs for internal combustion engines, and more specifically for more efficient novel variable valve actuation switching rocker arm systems. 
     BACKGROUND 
     Global environmental and economic concerns regarding increasing fuel consumption and greenhouse gas emission, the rising cost of energy worldwide, and demands for lower operating cost, are driving changes to legislative regulations and consumer demand. As these regulations and requirements become more stringent, advanced engine technologies must be developed and implemented to realize desired benefits. 
       FIG. 1B  illustrates several valve train arrangements in use today. In both Type I ( 21 ) and Type II ( 22 ), arrangements, a cam shaft with one or more valve actuating lobes  30  is located above an engine valve  29  (overhead cam). In a Type I ( 21 ) valvetrain, the overhead cam lobe  30  directly drives the valve through a hydraulic lash adjuster (HLA)  812 . In a Type II ( 22 ) valve train, an overhead cam lobe  30  drives a rocker arm  25 , and the first end of the rocker arm pivots over an HLA  812 , while the second end actuates the valve  29 . 
     In Type III ( 23 ), the first end of the rocker arm  28  rides on and is positioned above a cam lobe  30  while the second end of the rocker arm  28  actuates the valve  29 . As the cam lobe  30  rotates, the rocker arm pivots about a fixed shaft  31 . An HLA  812  can be implemented between the valve  29  tip and the rocker arm  28 . 
     In Type V ( 24 ), the cam lobe  30  indirectly drives the first end of the rocker arm  26  with a push rod  27 . An HLA  812  is shown implemented between the cam lobe  30  and the push rod  27 . The second end of the rocker arm  26  actuates the valve  29 . As the cam lobe  30  rotates, the rocker arm pivots about a fixed shaft  31 . 
     As  FIG. 1A  also illustrates, industry projections for Type II ( 22 ) valve trains in automotive engines, shown as a percentage of the overall market, are predicted to be the most common configuration produced by 2019. 
     Technologies focused on Type II ( 22 ) valve trains, that improve the overall efficiency of the gasoline engine by reducing friction, pumping, and thermal losses are being introduced to make the best use of the fuel within the engine. Some of these variable valve actuation (VVA) technologies have been introduced and documented. 
     A VVA device may be a variable valve lift (VVL) system, a cylinder deactivation (CDA) system such as that described U.S. patent application Ser. No. 13/532,777, filed Jun. 25, 2012 “Single Lobe Deactivating Rocker Arm,” hereby incorporated by reference in its entirety, or other valve actuation system. As noted, these mechanisms are developed to improve performance, fuel economy, and/or reduce emissions of the engine. Several types of the VVA rocker arm assemblies include an inner rocker arm within an outer rocker arm that are biased together with torsion springs. A latch, when in the latched position causes both the inner and outer rocker arms to move as a single unit. When unlatched, the rocker arms are allowed to move independent of each other. 
     Switching rocker arms allow for control of valve actuation by alternating between latched and unlatched states, usually involving the inner arm and outer arm, as described above. In some circumstances, these arms engage different cam lobes, such as low-lift lobes, high-lift lobes, and no-lift lobes. Mechanisms are required for switching rocker arm modes in a manner suited for operation of internal combustion engines. Rocker arms that are driven by a camshaft to actuate the cylinder intake and exhaust valves are typically mounted on the cylinder head. There are structures extending from the cylinder head such as cam towers to secure the camshafts in an overhead cam design. There are also spark plug tubes that extend upward from the top of each cylinder through the head to receive spark plugs. 
     As described above, some embodiments of VVA switching rocker arm assemblies include a rocker arm within a rocker arm that are biased together with a spring on either side. Since the inner/outer arm design often employs a roller in the center to engage a cam lobe, it is advantageous to keep the roller the same width of the cam lobe. Therefore, the structures on either side of the roller add width to the rocker assembly causing it to be wider than original non-VVA rocker arms and too wide to fit certain cylinder head designs. 
     For example, some Type II engine heads employ cam towers that have a hydraulic lifter adjuster (HLA) near the centerline of the head and spark plug tubes that obstruct one side of a wide VVA switching rocker arm assembly. 
     Many engine parts are designed by manufacturers to work with a specific cylinder head, making the cylinder head very difficult to modify because changes may impact many interrelated components, possibly increasing cost and causing assembly clearance issues. 
     One example of VVA technology used to alter operation and improve fuel economy in Type II gasoline engines is discrete variable valve lift (DVVL), also sometimes referred to as a DVVL switching rocker arm. DVVL works by limiting engine cylinder intake air flow with an engine valve that uses discrete valve lift states versus standard “part throttling”. A second example is cylinder deactivation (CDA). Fuel economy can be improved by using CDA at partial load conditions in order to operate select combustion cylinders at higher loads while turning off other cylinders. 
     The United States Environmental Protection Agency (EPA) showed a 4% improvement in fuel economy when using DVVL applied to various passenger car engines. An earlier report, sponsored by the United States Department of Energy lists the benefit of DVVL at 4.5% fuel economy improvement. Since automobiles spend most of their life at “part throttle” during normal cruising operation, a substantial fuel economy improvement can be realized when these throttling losses are minimized. For CDA, studies show a fuel economy gain, after considering the minor loss due to the deactivated cylinders, ranging between 2 and 14%. Currently, there is a need for VVA rocker arms for increased performance, economy and/or reduced emissions that fit specific engine head designs. 
     SUMMARY 
     Advanced VVA systems for piston-type internal combustion engines combine valve lift control devices, such as CDA or DVVL switching rocker arms, valve lift actuation methods, such as hydraulic actuation using pressurized engine oil, software and hardware control systems, and enabling technologies. Enabling technologies may include sensing and instrumentation, OCV design, DFHLA design, torsion springs, specialized coatings, algorithms, physical arrangements, etc. 
     In an embodiment, a modified rocker assembly is disclosed having an obstructed side and a non-obstructed side, having an outer structure having a first end, an inner rocker structure fitting within the outer structure, the inner structure also having a first end. The modified rocker assembly has an axle pivotally connecting the first ends of inner structure to the outer structure, such that the inner structure may rotate within the outer structure around the axle. At least one torsion spring on one side of axle, rotationally biases the inner structure relative to the outer structure. The outer structure, on the obstructed side as it extends from the second end toward the first end is offset toward the non-obstructed side creating a first offset portion to provide additional clearance on the obstructed side. This design allows the modified rocker arm to fit into an engine head having an obstruction on its obstruction side. 
     In an embodiment, a modified rocker assembly is disclosed having an obstructed side and a non-obstructed side, with an outer structure having a first end, an inner rocker structure fitting within the outer structure, the inner structure also having a first end. An axle pivotally connects the first ends of inner structure to the outer structure, such that the inner structure may rotate within the outer structure around the axle. At least one torsion spring is mounted on the non-obstructed side of the axle that rotationally biases the inner structure relative to the outer structure. As the outer structure on the obstructed side extends from the second end toward the first end, the outer structure is offset toward the non-obstructed side creating a first offset portion. The first offset portion provides additional clearance on the obstructed side. 
     In an embodiment, a modified rocker assembly is disclosed having an obstructed side and a non-obstructed side. The modified rocker assembly has an outer structure having a first end with an offset portion, an inner rocker structure fitting within the outer structure. The inner structure also has a first end. An axle pivotally connects the first ends of inner structure to the outer structure, such that the inner structure may rotate within the outer structure around the axle. The modified rocker assembly has at least one torsion spring on one side of the axle, rotationally biasing the inner structure relative to the outer structure. As the outer structure on the obstructed side extends from the second end toward the first end, the outer structure smoothly curves toward the non-obstructed side. This creates a first offset portion that provides additional clearance on the obstructed side. This allows this embodiment to fit in an engine head that has an obstruction on the obstructed side. 
     In one embodiment, an advanced discrete variable valve lift (DVVL) system is described. The advanced discrete variable valve lift (DVVL) system was designed to provide two discrete valve lift states in a single rocker arm. Embodiments of the approach presented relate to the Type II valve train described above and shown in  FIG. 1B . Embodiments of the system presented herein may apply to a passenger car engine (having four cylinders in embodiments) with an electro-hydraulic oil control valve, dual feed hydraulic lash adjuster (DFHLA), and DVVL switching rocker arm. The DVVL switching rocker arm embodiments described herein focus on the design and development of a switching roller finger follower (SRFF) rocker arm system which enables two-mode discrete variable valve lift on end pivot roller finger follower valve trains. This switching rocker arm configuration includes a low friction roller bearing interface for the low lift event, and retains normal hydraulic lash adjustment for maintenance free valve train operation. 
     Mode switching (i.e., from low to high lift or vice versa) is accomplished within one cam revolution, resulting in transparency to the driver. The SRFF prevents significant changes to the overhead required for installing in existing engine designs. Load carrying surfaces at the cam interface may comprise a roller bearing for low lift operation, and a diamond like carbon coated slider pad for high lift operation. Among other aspects, the teachings of the present application is able to reduce mass and moment of inertia while increasing stiffness to achieve desired dynamic performance in low and high lift modes. 
     A diamond-like carbon coating (DLC coating) allows higher slider interface stresses in a compact package. Testing results show that this technology is robust and meets all lifetime requirements with some aspects extending to six times the useful life requirements. Alternative materials and surface preparation methods were screened, and results showed DLC coating to be the most viable alternative. This application addresses the technology developed to utilize a Diamond-like carbon (DLC) coating on the slider pads of the DVVL switching rocker arm. 
     System validation test results reveal that the system meets dynamic and durability requirements. Among other aspects, this patent application also addresses the durability of the SRFF design for meeting passenger car durability requirements. Extensive durability tests were conducted for high speed, low speed, switching, and cold start operation. High engine speed test results show stable valve train dynamics above 7000 engine rpm. System wear requirements met end-of-life criteria for the switching, sliding, rolling and torsion spring interfaces. One important metric for evaluating wear is to monitor the change in valve lash. The lifetime requirements for wear showed that lash changes are within the acceptable window. The mechanical aspects exhibited robust behavior over all tests including the slider interfaces that contain a diamond like carbon (DLC) coating. 
     With flexible and compact packaging, this DVVL system can be implemented in a multi-cylinder engine. The DVVL arrangement can be applied to any combination of intake or exhaust valves on a piston-driven internal combustion engine. Enabling technologies include OCV, DFHLA, DLC coating. 
     In a second embodiment, an advanced single-lobe cylinder deactivation (CDA-1L) system is described. The advanced cylinder deactivation (CDA-1L) system was designed to deactivate one or more cylinders. Embodiments of the approach presented relate to the Type II valve train described above and shown in  FIG. 22 . Embodiments of the system presented herein may apply to a passenger car engine (having a multiple of two cylinders in embodiments, for example 2, 6, 8) with an electro-hydraulic oil control valve, dual feed hydraulic lash adjuster (DFHLA), and CDA-1L switching rocker arm. The CDA-1L switching rocker arm embodiments described herein focus on the design and development of a switching roller finger follower (SRFF) rocker arm system which enables lift/no-lift operation for end pivot roller finger follower valve trains. This switching rocker arm configuration includes a low friction roller bearing interface for the cylinder deactivation event, and retains normal hydraulic lash adjustment for maintenance free valve train operation. 
     Mode switching for the CDA-1L system is accomplished within one cam revolution, resulting in transparency to the driver. The SRFF prevents significant changes to the overhead required for installing in existing engine designs. Among other aspects, the teachings of the present application is able to reduce mass and moment of inertia while increasing stiffness to achieve desired dynamic performance in either lift or no-lift modes. 
     CDA-1L system validation test results reveal that the system meets dynamic and durability requirements. Among other aspects, this patent application also addresses the durability of the SRFF design necessary to meet passenger car durability requirements. Extensive durability tests were conducted for high speed, low speed, switching, and cold start operation. High engine speed test results show stable valve train dynamics above 7000 engine rpm. System wear requirements met end-of-life criteria for the switching, rolling and torsion spring interfaces. One important metric for evaluating wear is to monitor the change in valve lash. The lifetime requirements for wear showed that lash changes are within the acceptable window. The mechanical aspects exhibited robust behavior over all tests. 
     With flexible and compact packaging, the CDA-1L system can be implemented in a multi-cylinder engine. Enabling technologies include OCV, DFHLA, and specialized torsion spring design. 
     A rocker arm is described for engaging a cam having one lift lobe per valve. The rocker arm includes an outer arm, an inner arm, a pivot axle, a lift lobe contacting bearing, a bearing axle, and at least one bearing axle spring. The outer arm has a first and a second outer side arms and outer pivot axle apertures configured for mounting the pivot axle. The inner arm is disposed between the first and second outer side arms, and has a first inner side arm and a second inner side arm. The first and second inner side arms have an inner pivot axle apertures that receive and hold the pivot axle, and inner bearing axle apertures for mounting the bearing axle. 
     The pivot axle fits into the inner pivot axle apertures and the outer pivot axle apertures. 
     The bearing axle is mounted in the bearing axle apertures of the inner arm. 
     The bearing axle spring is secured to the outer arm and is in biasing contact with the bearing axle. The lift lobe contacting bearing is mounted to the bearing axle between the first and the second inner side arms. 
     Another embodiment can be described as a rocker arm for engaging a cam having a single lift lobe per engine valve. The rocker arm includes an outer arm, an inner arm, a cam contacting member configured to be capable of transferring motion from the single lift lobe of the cam to the rocker arm, and at least one biasing spring. 
     The rocker arm also includes a first outer side arm and a second outer side arm. 
     The inner arm is disposed between the first and the second outer side arms, and has a first inner side arm and a second inner side arm. 
     The inner arm is secured to the outer arm by a pivot axle configured to permit rotating movement of the inner arm relative to the outer arm about the pivot axle. 
     The cam contacting member is disposed between the first and second inner side arm. 
     At least one biasing spring is secured to the outer arm and is in biasing contact with the cam contacting member. 
     Another embodiment may be described as a deactivating rocker arm for engaging a cam having a single lift lobe having a first end and a second end, an outer arm, an inner arm, a pivot axle, a lift lobe contacting member configured to be capable of transferring motion from the cam lift lobe to the rocker arm, a latch configured to be capable of selectively deactivating the rocker arm, and at least one biasing spring. 
     The outer arm has a first outer side arm and a second outer side arm, outer pivot axle apertures configured for mounting the pivot axle, and axle slots configured to accept the lift lobe contacting member, permitting lost motion movement of the lift lobe contacting member. 
     The inner arm is disposed between the first and second outer side arms, and has a first inner side arm and a second inner side arm. The first inner side arm and the second inner side arm have inner pivot axle apertures configured for mounting the pivot axle, and inner lift lobe contacting member apertures configured for mounting the lift lobe contacting member. 
     The pivot axle is mounted adjacent the first end of the rocker arm and disposed in the inner pivot axle apertures and the outer pivot axle apertures. 
     The latch is disposed adjacent the second end of the rocker arm. 
     The lift lobe contacting member mounted in the lift lobe contacting member apertures of the inner arm and the axle slots of the outer arm and between the pivot axle and latch. 
     The biasing spring is secured to the outer arm and in biasing contact with the lift lobe contacting member. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       It will be appreciated that the illustrated boundaries of elements in the drawings represent only one example of the boundaries. One of ordinary skill in the art will appreciate that a single element may be designed as multiple elements or that multiple elements may be designed as a single element. An element shown as an internal feature may be implemented as an external feature and vice versa. 
       Further, in the accompanying drawings and description that follow, like parts are indicated throughout the drawings and description with the same reference numerals, respectively. The figures may not be drawn to scale and the proportions of certain parts have been exaggerated for convenience of illustration. 
         FIG. 1A  illustrates the relative percentage of engine types for 2012 and 2019. 
         FIG. 1B  illustrates the general arrangement and market sizes for Type I, Type II, Type III, and Type V valve trains. 
         FIG. 2  shows the intake and exhaust valve train arrangement. 
         FIG. 3  illustrates the major components that comprise the DVVL system, including hydraulic actuation. 
         FIG. 4  illustrates a perspective view of an exemplary switching rocker arm as it may be configured during operation with a three lobed cam. 
         FIG. 5  is a diagram showing valve lift states plotted against cam shaft crank degrees for both the intake and exhaust valves for an exemplary DVVL implementation. 
         FIG. 6  is a system control diagram for a hydraulically actuated DVVL rocker arm assembly. 
         FIG. 7  illustrates the rocker arm oil gallery and control valve arrangement. 
         FIG. 8  illustrates the hydraulic actuating system and conditions for an exemplary DVVL switching rocker arm system during low-lift (unlatched) operation. 
         FIG. 9  illustrates the hydraulic actuating system and conditions for an exemplary DVVL switching rocker arm system during high-lift (latched) operation. 
         FIG. 10  illustrates a side cut-away view of an exemplary switching rocker arm assembly with dual feed hydraulic lash adjuster (DFHLA). 
         FIG. 11  is a cut-away view of a DFHLA. 
         FIG. 12  illustrates diamond like carbon coating layers. 
         FIG. 13  illustrates an instrument used to sense position or relative movement of a DFHLA ball plunger. 
         FIG. 14  illustrates an instrument used in conjunction with a valve stem to measure valve movement relative to a known state. 
         FIGS. 14A and 14B  illustrate a section view of a first linear variable differential transformer using three windings to measure valve stem movement. 
         FIGS. 14C and 14D  illustrate a section view of a second linear variable differential transformer using two windings to measure valve stem movement. 
         FIG. 15  illustrates another perspective view of an exemplary switching rocker arm and  FIG. 15A  depicts the exemplary switching rocker arm held in a clamping fixture. 
         FIG. 16  illustrates an instrument designed to sense position and/or movement. 
         FIG. 17  is a graph that illustrates the relationship between OCV actuating current, actuating oil pressure, and valve lift state during a transition between high-lift and low-lift states. 
         FIG. 17A  is a graph that illustrates the relationship between OCV actuating current, actuating oil pressure, and latch state during a latch transition. 
         FIG. 17B  is a graph that illustrates the relationship between OCV actuating current, actuating oil pressure, and latch state during another latch transition. 
         FIG. 17C  is a graph that illustrates the relationship between valve lift profiles and actuating oil pressure for high-lift and low-lift states. 
         FIG. 18  is a control logic diagram for a DVVL system. 
         FIG. 19  illustrates an exploded view of an exemplary switching rocker arm. 
         FIG. 20  is a chart illustrating oil pressure conditions and oil control valve (OCV) states for both low-lift and high-lift operation of a DVVL rocker arm assembly. 
         FIGS. 21-22  illustrate graphs showing the relation between oil temperature and latch response time. 
         FIG. 23  is a timing diagram showing available switching windows for an exemplary DVVL switching rocker arm, in a 4-cylinder engine, with actuating oil pressure controlled by two OCV&#39;s each controlling two cylinders. 
         FIG. 24  is a side cutaway view of a DVVL switching rocker arm illustrating latch pre-loading prior to switching from high-lift to low-lift. 
         FIG. 25  is a side cutaway view of a DVVL switching rocker arm illustrating latch pre-loading prior to switching from low-lift to high-lift. 
         FIG. 25A  is a side cutaway view of a DVVL switching rocker arm illustrating a critical shift event when switching between low-lift and high-lift. 
         FIG. 26  is an expanded timing diagram showing available switching windows and constituent mechanical switching times for an exemplary DVVL switching rocker arm, in a 4-cylinder engine, with actuating oil pressure controlled by two OCV&#39;s each controlling two cylinders. 
         FIG. 27  illustrates a perspective view of an exemplary switching rocker arm. 
         FIG. 28  illustrates a top-down view of exemplary switching rocker arm. 
         FIG. 29  illustrates a cross-section view taken along line  29 - 29  in  FIG. 28 . 
         FIGS. 30A-30B  illustrate a section view of an exemplary torsion spring. 
         FIG. 31  illustrates a bottom perspective view of the outer arm. 
         FIG. 32  illustrates a cross-sectional view of the latching mechanism in its latched state along the line  32 ,  33 - 32 ,  33  in  FIG. 28 . 
         FIG. 33  illustrates a cross-sectional view of the latching mechanism in its unlatched state. 
         FIG. 34  illustrates an alternate latch pin design. 
         FIGS. 35A-35F  illustrate several retention devices for orientation pin. 
         FIG. 36  illustrates an exemplary latch pin design. 
         FIG. 37  illustrates an alternative latching mechanism. 
         FIGS. 38-40  illustrate an exemplary method of assembling a switching rocker arm. 
         FIG. 41  illustrates an alternative embodiment of pin. 
         FIG. 42  illustrates an alternative embodiment of a pin. 
         FIG. 43  illustrates the various lash measurements of a switching rocker arm. 
         FIG. 44  illustrates a perspective view of an exemplary inner arm of a switching rocker arm. 
         FIG. 45  illustrates a perspective view from below of the inner arm of a switching rocker arm. 
         FIG. 46  illustrates a perspective view of an exemplary outer arm of a switching rocker arm. 
         FIG. 47  illustrates a sectional view of a latch assembly of an exemplary switching rocker arm. 
         FIG. 48  is a graph of lash vs. camshaft angle for a switching rocker arm. 
         FIG. 49  illustrates a side cut-away view of an exemplary switching rocker arm assembly. 
         FIG. 50  illustrates a perspective view of the outer arm with an identified region of maximum deflection when under load conditions. 
         FIG. 51  illustrates a top view of an exemplary switching rocker arm and three-lobed cam. 
         FIG. 52  illustrates a section view along line  52 - 52  in of  FIG. 51  of an exemplary switching rocker arm. 
         FIG. 53  illustrates an exploded view of an exemplary switching rocker arm, showing the major components that affect inertia for an exemplary switching rocker arm assembly. 
         FIG. 54  illustrates a design process to optimize the relationship between inertia and stiffness for an exemplary switching rocker assembly. 
         FIG. 55  illustrates a characteristic plot of inertia versus stiffness for design iterations of an exemplary switching rocker arm assembly. 
         FIG. 56  illustrates a characteristic plot showing stress, deflection, loading, and stiffness versus location for an exemplary switching rocker arm assembly. 
         FIG. 57  illustrates a characteristic plot showing stiffness versus inertia for a range of exemplary switching rocker arm assemblies. 
         FIG. 58  illustrates an acceptable range of discrete values of stiffness and inertia for component parts of multiple DVVL switching rocker arm assemblies. 
         FIG. 59  is a side cut-away view of an exemplary switching rocker arm assembly including a DFHLA and valve. 
         FIG. 60  illustrates a characteristic plot showing a range of stiffness values versus location for component parts of an exemplary switching rocker arm assembly. 
         FIG. 61  illustrates a characteristic plot showing a range of mass distribution values versus location for component parts of an exemplary switching rocker arm assembly. 
         FIG. 62  illustrates a test stand measuring latch displacement. 
         FIG. 63  is an illustration of a non-firing test stand for testing switching rocker arm assembly. 
         FIG. 64  is a graph of valve displacement vs. camshaft angle. 
         FIG. 65  illustrates a hierarchy of key tests for testing the durability of a switching roller finger follower (SRFF) rocker arm assembly. 
         FIG. 66  shows the test protocol in evaluating the SRFF over an Accelerated System Aging test cycle. 
         FIG. 67  is a pie chart showing the relative testing time for the SRFF durability testing. 
         FIG. 68  shows a strain gage that was attached to and monitored the SRFF during testing. 
         FIG. 69  is a graph of valve closing velocity for the Low Lift mode. 
         FIG. 70  is a valve drop height distribution. 
         FIG. 71  displays the distribution of critical shifts with respect to camshaft angle. 
         FIG. 72  show an end of a new outer arm before use. 
         FIG. 73  shows typical wear of the outer arm after use. 
         FIG. 74  illustrates average Torsion Spring Load Loss at end-of-life testing. 
         FIG. 75  illustrates the total mechanical lash change of Accelerated System Aging Tests. 
         FIG. 76  illustrates end-of-life slider pads with the DLC coating, exhibiting minimal wear. 
         FIG. 77  is a camshaft surface embodiment employing a crown shape. 
         FIG. 78  illustrates a pair of slider pads attached to a support rocker on a test coupon. 
         FIG. 79A  illustrates DLC coating loss early in the testing of a coupon. 
         FIG. 79B  shows a typical example of one of the coupons tested at the max design load with 0.2 degrees of included angle. 
         FIG. 80  is a graph of tested stress level vs. engine lives for a test coupon having DLC coating. 
         FIG. 81  is a graph showing the increase in engine lifetimes for slider pads having polished and non-polished surfaces prior to coating with a DLC coating. 
         FIG. 82  is a flowchart illustrating the development of the production grinding and polishing processes that took place concurrently with the testing. 
         FIG. 83  shows the results of the slider pad angle control relative to three different grinders. 
         FIG. 84  illustrates surface finish measurements for three different grinders. 
         FIG. 85  illustrates the results of six different fixtures to hold the outer arm during the slider pad grinding operations. 
         FIG. 86  is a graph of valve closing velocity for the High Lift mode. 
         FIG. 87  illustrates durability test periods. 
         FIG. 88  shows a perspective view of an exemplary CDA-1L layout. 
         FIG. 89A  shows a partial cut-away side elevational view of an exemplary SRFF-1L system with a latch mechanism and roller bearing. 
         FIG. 89B  shows a front elevation view of the exemplary SRFF-1L system of  FIG. 89A . 
         FIG. 90  is an engine layout showing an exemplary SRFF-1L rocker assembly on the exhaust and intake valves. 
         FIG. 91  shows a hydraulic fluid control system. 
         FIG. 92  shows an exemplary SRFF-1L system in operation exhibiting normal-lift engine valve operation. 
         FIGS. 93A, 93B and 93C  show an exemplary SRFF-1L system in operation exhibiting no-lift engine valve operation. 
         FIG. 94  shows an example switching window. 
         FIG. 95  shows the effect of camshaft phasing on the switching window. 
         FIG. 96  shows latch response times for an embodiment of the SRFF-1 system. 
         FIG. 97  is a graph showing a switching window times above 40 degrees C. for an exemplary SRFF-1 system. 
         FIG. 98  is a graph showing a switching window times taking into account camshaft phasing and oil temperature for an exemplary SRFF-1 system. 
         FIG. 99  illustrates an exemplary SRFF-1L rocker arm assembly. 
         FIG. 100  illustrates an exploded view of the exemplary SRFF-1L rocker arm assembly of  FIG. 99 . 
         FIG. 101  illustrates a side view of an exemplary SRFF-1L rocker arm assembly, including DFHLA, valve stem and cam lobe. 
         FIG. 102  illustrates an end view of an exemplary SRFF-1L rocker arm assembly, including DFHLA, valve stem and cam lobe. 
         FIG. 103  shows latch re-engagement features in case of pressure loss. 
         FIG. 104  shows camshaft alignment of an exemplary SRFF-1L system. 
         FIG. 105  shows forces acting on an RFF employing hydraulic lash adjusters. 
         FIG. 106  shows a force balance for an exemplary SRFF-1L system in a ‘no-lift’ mode. 
         FIG. 107  is a table showing oil pressure requirements for an exemplary SRFF-1 system. 
         FIG. 108  shows mechanical lash for an exemplary SRFF-1 system. 
         FIG. 109  shows camshaft lift profiles for a three-lobe CDA system versus an exemplary SRFF-1L system. 
         FIG. 110  is a graphic representation of stiffness vs. moment of inertia for multiple rocker arm designs. 
         FIG. 111  illustrates the resultant seating closing velocity of an intake valve of an exemplary SRFF-1L system. 
         FIG. 112  is a table showing a torsion spring test summary. 
         FIG. 113  is a graph showing displacements and pressures during a ‘pump-up’ test. 
         FIG. 114  shows durability and lash change over a specified testing period for an exemplary STFF-1L system. 
         FIG. 115  is a perspective view of a prior art cylinder head with parts removed for clarity. 
         FIG. 116  is an elevational, cross-sectional view of the cylinder head of  FIG. 115 . 
         FIG. 117  is a perspective view of a prior art variable valve lift (VVL) rocker arm assembly. 
         FIG. 118  is a perspective view of a left-handed (modified) rocker assembly that provides variable valve lift according to one aspect of the present teachings. 
         FIG. 119  is a top plan view of the modified rocker assembly of  FIG. 118 . 
         FIG. 120  is a side elevational view of the modified rocker assembly  400  of  FIGS. 118-119 . 
         FIG. 121  is an end-on elevational view of the modified rocker assembly of  FIGS. 118-120  as viewed from its hinge (first) end. 
         FIG. 122  is an end-on elevational view of the modified rocker assembly of  FIGS. 118-121  as viewed from its latch (second) end. 
         FIG. 123  is a plan view from above the outer structure showing the first and second offset areas. 
         FIG. 124  is a plan view from below the outer structure of  FIG. 123 . 
         FIG. 125  is a side elevational view of an outer structure according to one aspect of the present teachings. 
         FIG. 126  is a perspective view of top side of an inner structure according to one aspect of the present teachings. 
         FIG. 127  is a perspective view of bottom side of the inner structure of  FIG. 126 . 
         FIG. 128  is a plan view from the top side of the inner structure of  FIGS. 126-127 . 
         FIG. 129  is a plan view from the bottom side of the inner structure of  FIGS. 126-128 . 
         FIG. 130  is an end-on elevational view of the inner structure of  FIGS. 126-129  as viewed from its hinge (first) end. 
         FIG. 131  is an end-on elevational view of the inner structure of  FIGS. 126-130  as viewed from its latch (second) end. 
         FIG. 132  is a perspective view of the modified rocker assembly of  FIGS. 118-122  as it would appear installed in a cylinder head. 
         FIG. 133  is a perspective view from another viewpoint of the modified rocker assembly  400  of  FIGS. 118-122 , as it would appear installed in a cylinder head. 
     
    
    
     DETAILED DESCRIPTION 
     The terms used herein have their common and ordinary meanings unless redefined in this specification, in which case the new definitions will supersede the common meanings. 
     It is also to be appreciated that the phraseology and terminology used herein is for the purpose of description and should not be regarded as limiting. References in the singular or plural form are not intended to limit the presently disclosed systems or methods, their components, acts, or elements. The use herein of “including,” “comprising,” “having,” “containing,” “involving,” and variations thereof is meant to encompass the items listed thereafter and equivalents thereof as well as additional items. References to “or” may be construed as inclusive so that any terms described using “or” may indicate any of a single, more than one, and all of the described terms. Any references to front and back, left and right, top and bottom, and upper and lower are intended for convenience of description, not to limit the present systems and methods or their components to any one positional or spatial orientation. 
     As illustrated in the various figures, some sizes of structures or portions are exaggerated relative to other structures or portions for illustrative purposes and, thus, are provided to illustrate the general structures of the present subject matter. Furthermore, various aspects of the present subject matter are described with reference to a structure or a portion being formed on other structures, portions, or both. As will be appreciated by those of skill in the art, references to a structure being formed “on” or “above” another structure or portion contemplates that additional structure, portion, or both may intervene. References to a structure or a portion being formed “on” another structure or portion without an intervening structure or portion are described herein as being formed “directly on” the structure or portion. Similarly, it will be understood that when an element is referred to as being “connected”, “attached”, or “coupled” to another element, it can be directly connected, attached, or coupled to the other element, or intervening elements may be present. In contrast, when an element is referred to as being “directly connected”, “directly attached”, or “directly coupled” to another element, no intervening elements are present. 
     Furthermore, relative terms such as “on”, “above”, “upper”, “top”, “lower”, or “bottom” are used herein to describe one structure&#39;s or portion&#39;s relationship to another structure or portion as illustrated in the figures. It will be understood that relative terms such as “on”, “above”, “upper”, “top”, “lower” or “bottom” are intended to encompass different orientations of the device in addition to the orientation depicted in the figures. For example, if the device in the figures is turned over, structure or portion described as “above” other structures or portions would now be oriented “below” the other structures or portions. Likewise, if devices in the figures are rotated along an axis, structure or portion described as “above”, other structures or portions would now be oriented “next to” or “left of” the other structures or portions. Like numbers refer to like elements throughout. 
     VVA SYSTEM EMBODIMENTS—VVA system embodiments represent a unique combination of a switching device, actuation method, analysis and control system, and enabling technology that together produce a VVA system. VVA system embodiments may incorporate one or more enabling technologies. 
     I. Discrete Variable Valve Lift (DVVL) System Embodiment Description 
     1. DVVL System Overview 
     A cam-driven, discrete variable valve lift (DVVL), switching rocker arm device that is hydraulically actuated using a combination of dual-feed hydraulic lash adjusters (DFHLA), and oil control valves (OCV) is described in following sections as it would be installed on an intake valve in a Type II valve train. In alternate embodiments, this arrangement can be applied to any combination of intake or exhaust valves on a piston-driven internal combustion engine. 
     As illustrated in  FIG. 2 , the exhaust valve train in this embodiment comprises a fixed rocker arm  810 , single lobe camshaft  811 , a standard hydraulic lash adjuster (HLA)  812 , and an exhaust valve  813 . As shown in  FIGS. 2 and 3 , components of the intake valve train include the three-lobe camshaft  102 , switching rocker arm assembly  100 , a dual feed hydraulic lash adjuster (DFHLA)  110  with an upper fluid port  506  and a lower fluid port  512 , and an electro-hydraulic solenoid oil control valve assembly (OCV)  820 . The OCV  820  has an inlet port  821 , and a first and second control port  822 ,  823  respectively. 
     Referring to  FIG. 2 , the intake and exhaust valve trains share certain common geometries including valve  813  spacing to HLA  812  and valve  112  spacing to DFHLA  110 . Maintaining a common geometry allows the DVVL system to package with existing or lightly modified Type II cylinder head space while utilizing the standard chain drive system. Additional components, illustrated in  FIG. 4 , that are common to both the intake and exhaust valve train include valves  112 , valve springs  114 , and valve spring retainers  116 . Valve keys and valve stem seals (not shown) are also common for both the intake and exhaust. Implementation cost for the DVVL system is minimized by maintaining common geometries, using common components. 
     The intake valve train elements illustrated in  FIG. 3  work in concert to open the intake valve  112  with either high-lift camshaft lobes  104 ,  106  or a low-lift camshaft lobe  108 . The high-lift camshaft lobes  104 ,  106  are designed to provide performance comparable to a fixed intake valve train, and are comprised of a generally circular portion where no lift occurs, a lift portion, which may include a linear lift transition portion, and a nose portion that corresponds to maximum lift. The low-lift camshaft lobe  108  allows for lower valve lift and early intake valve closing. The low-lift camshaft lobe  108  also comprises a generally circular portion where no lift occurs, a generally linear portion were lift transitions, and a nose portion that corresponds to maximum lift. The graph in  FIG. 5  shows a plot of valve lift  818  versus crank angle  817 . The cam shaft high-lift profile  814  and the fixed exhaust valve lift profile  815  are contrasted with low-lift profile  816 . The low-lift event illustrated by profile  816  reduces both lift and duration of the intake event during part throttle operation to decrease throttling losses and realize a fuel economy improvement. This is also referred to as early intake valve closing, or EIVC. When full power operation is needed, the DVVL system returns to the high-lift profile  814 , which is similar to a standard fixed lift event. Transitioning from low-lift to high-lift and vice versa occurs within one camshaft revolution. The exhaust lift event shown by profile  815  is fixed and operates in the same way with either a low-lift or high-lift intake event. 
     The system used to control DVVL switching uses hydraulic actuation. A schematic depiction of a hydraulic control and actuation system  800  that is used with embodiments of the teachings of the present application is shown in  FIG. 6 . The hydraulic control and actuation system  800  is designed to deliver hydraulic fluid, as commanded by controlled logic, to mechanical latch assemblies that provide for switching between high-lift and low-lift states. An engine control unit  825  controls when the mechanical switching process is initiated. The hydraulic control and actuation system  800  shown is for use in a four cylinder in-line Type II engine on the intake valve train described previously, though the skilled artisan will appreciate that control and actuation system may apply to engines of other “Types” and different numbers of cylinders. 
     Several enabling technologies previously mentioned and used in the DVVL system described herein may be used in combination with other DVVL system components described herein thus rending unique combinations, some of which will be described herein: 
     2. DVVL System Enabling Technologies 
     Several technologies used in this system have multiple uses in varied applications; they are described herein as components of the DVVL system disclosed herein. These include: 
     2.1. Oil Control Valve (OCV) and Oil Control Valve Assemblies 
     Now, referring to  FIGS. 7-9 , an OCV is a control device that directs or does not direct pressurized hydraulic fluid to cause the rocker arm  100  to switch between high-lift mode and low-lift mode. OCV activation and deactivation is caused by a control device signal  866 . One or more OCVs can be packaged in a single module to form an assembly. In one embodiment, OCV assembly  820  is comprised of two solenoid type OCV&#39;s packaged together. In this embodiment, a control device provides a signal  866  to the OCV assembly  820 , causing it to provide a high pressure (in embodiments, at least 2 Bar of oil pressure) or low pressure (in embodiments, 0.2-0.4 Bar) oil to the oil control galleries  802 ,  803  causing the switching rocker arm  100  to be in either low-lift or high-lift mode, as illustrated in  FIGS. 8 and 9  respectively. Further description of this OCV assembly  820  embodiment is contained in following sections. 
     2.2. Dual Feed Hydraulic Lash Adjuster (DFHLA): 
     Many hydraulic lash adjusting devices exist for maintaining lash in engines. For DVVL switching of rocker arm  100  ( FIG. 4 ), traditional lash management is required, but traditional HLA devices are insufficient to provide the necessary oil flow requirements for switching, withstand the associated side-loading applied by the assembly  100  during operation, and fit into restricted package spaces. A compact dual feed hydraulic lash adjuster  110  (DFHLA), used together with a switching rocker arm  100  is described, with a set of parameters and geometry designed to provide optimized oil flow pressure with low consumption, and a set of parameters and geometry designed to manage side loading. 
     As illustrated in  FIG. 10 , the ball plunger end  601  fits into the ball socket  502  that allows rotational freedom of movement in all directions. This permits side and possibly asymmetrical loading of the ball plunger end  601  in certain operating modes, for example when switching from high-lift to low-lift and vice versa. In contrast to typical ball end plungers for HLA devices, the DFHLA  110  ball end plunger  601  is constructed with thicker material to resist side loading, shown in  FIG. 11  as plunger thickness  510 . 
     Selected materials for the ball plunger end  601  may also have higher allowable kinetic stress loads, for example, chrome vanadium alloy. 
     Hydraulic flow pathways in the DFHLA  110  are designed for high flow and low pressure drop to ensure consistent hydraulic switching and reduced pumping losses. The DFHLA is installed in the engine in a cylindrical receiving socket sized to seal against exterior surface  511 , illustrated in  FIG. 11 . The cylindrical receiving socket combines with the first oil flow channel  504  to form a closed fluid pathway with a specified cross-sectional area. 
     As shown in  FIG. 11 , the preferred embodiment includes four oil flow ports  506  (only two shown) as they are arranged in an equally spaced fashion around the base of the first oil flow channel  504 . Additionally, two second oil flow channels  508  are arranged in an equally spaced fashion around ball end plunger  601 , and are in fluid communication with the first oil flow channel  504  through oil ports  506 . Oil flow ports  506  and the first oil flow channel  504  are sized with a specific area and spaced around the DFHLA  110  body to ensure even flow of oil and minimized pressure drop from the first flow channel  504  to the third oil flow channel  509 . The third oil flow channel  509  is sized for the combined oil flow from the multiple second oil flow channels  508 . 
     2.3. Diamond-Like Carbon Coating (DLC) 
     A diamond-like carbon coating (DLC) coating is described that can reduce friction between treated parts, and at the same provide necessary wear and loading characteristics Similar coating materials and processes exist, none are sufficient to meet many of the requirements encountered when used with VVA systems. For example, 1) be of sufficient hardness, 2) have suitable loadbearing capacity, 3) be chemically stable in the operating environment, 4) be applied in a process where temperatures do not exceed part annealing temperatures, 5) meet engine lifetime requirements, and 6) offer reduced friction as compared to a steel on steel interface. 
     A unique DLC coating process is described that meets the requirements set forth above. The DLC coating that was selected is derived from a hydrogenated amorphous carbon or similar material. The DLC coating is comprised of several layers described in  FIG. 12 .
         1. The first layer is a chrome adhesion layer  701  that acts as a bonding agent between the metal receiving surface  700  and the next layer  702 .   2. The second layer  702  is chrome nitride that adds ductility to the interface between the base metal receiving surface  700  and the DLC coating.   3. The third layer  703  is a combination of chrome carbide and hydrogenated amorphous carbon which bonds the DLC coating to the chrome nitride layer  702 .   4. The fourth layer  704  is comprised of hydrogenated amorphous carbon that provides the hard functional wear interface.       

     The combined thickness of layers  701 - 704  is between two and six micrometers. The DLC coating cannot be applied directly to the metal receiving surface  700 . To meet durability requirements and for proper adhesion of the first chrome adhesion layer  701  with the base receiving surface  700 , a very specific surface finish mechanically applied to the base layer receiving surface  700 . 
     2.4 Sensing and Measurement 
     Information gathered using sensors may be used to verify switching modes, identify error conditions, or provide information analyzed and used for switching logic and timing. Several sensing devices that may be used are described below. 
     2.4.1 Dual Feed Hydraulic Lash Adjuster (DFHLA) Movement 
     Variable valve actuation (VVA) technologies are designed to change valve lift profiles during engine operation using switching devices, for example a DVVL switching rocker arm or cylinder deactivation (CDA) rocker arm. When employing these devices, the status of valve lift is important information that confirms a successful switching operation, or detects an error condition/malfunction. 
     A DFHLA is used to both manage lash and supply hydraulic fluid for switching in VVA systems that employ switching rocker arm assemblies such as CDA or DVVL. As shown in the section view of  FIG. 10 , normal lash adjustment for the DVVL rocker arm assembly  100 , (a detailed description is in following sections) causes the ball plunger  601  to maintain contact with the inner arm  122  receiving socket during both high-lift and low-lift operation. The ball plunger  601  is designed to move as necessary when loads vary from between high-lift and low-lift states. A measurement of the movement  514  of  FIG. 13  in comparison with known states of operation can determine the latch location status. In one embodiment, a non-contact switch  513  is located between the HLA outer body and the ball plunger cylindrical body. A second example may incorporate a Hall-effect sensor mounted in a way that allows measurement of the changes in magnetic fields generated by a certain movement  514 . 
     2.4.2 Valve Stem Movement 
     Variable valve actuation (VVA) technologies are designed to change valve lift profiles during engine operation using switching devices, for example a DVVL switching rocker arm. The status of valve lift is important information that confirms a successful switching operation, or detects an error condition/malfunction. Valve stem position and relative movement sensors can be used to for this function. 
     One embodiment to monitor the state of VVA switching, and to determine if there is a switching malfunction is illustrated in  FIGS. 14 and 14A . In accordance with one aspect of the present teachings, a linear variable differential transformer (LVDT) type of transducer can convert the rectilinear motion of valve  872 , to which it is coupled mechanically, into a corresponding electrical signal. LVDT linear position sensors are readily available that can measure movements as small as a few millionths of an inch up to several inches. 
       FIG. 14A  shows the components of a typical LVDT installed in a valve stem guide  871 . The LVDT internal structure consists of a primary winding  899  centered between a pair of identically wound secondary windings  897 ,  898 . In embodiments, the windings  897 ,  898 ,  899  are wound in a recessed hollow formed in the valve guide body  871  that is bounded by a thin-walled section  878 , a first end wall  895 , and a second end wall  896 . In this embodiment, the valve guide body  871  is stationary. 
     Now, as to  FIGS. 14, 14A, and 14B , the moving element of this LVDT arrangement is a separate tubular armature of magnetically permeable material called the core  873 . In embodiments, the core  873  is fabricated into the valve  872  stem using any suitable method and manufacturing material, for example iron. 
     The core  873  is free to move axially inside the primary winding  899 , and secondary windings  897 ,  898 , and it is mechanically coupled to the valve  872 , whose position is being measured. There is no physical contact between the core  873 , and valve guide  871  inside bore. 
     In operation, the LVDT&#39;s primary winding,  899 , is energized by applying an alternating current of appropriate amplitude and frequency, known as the primary excitation. The magnetic flux thus developed is coupled by the core  873  to the adjacent secondary windings,  897  and  898 . 
     As shown in  14 A, if the core  873  is located midway between the secondary windings  897 ,  898 , an equal magnetic flux is then coupled to each secondary winding, making the respective voltages induced in windings  897  and  898  equal. At this reference midway core  873  position, known as the null point, the differential voltage output is essentially zero. 
     The core  873  is arranged so that it extends past both ends of winding  899 . As shown in  FIG. 14B , if the core  873  is moved a distance  870  to make it closer to winding  897  than to winding  898 , more magnetic flux is coupled to winding  897  and less to winding  898 , resulting in a non-zero differential voltage. Measuring the differential voltages in this manner can indicate both direction of movement and position of the valve  872 . 
     In a second embodiment, illustrated in  FIGS. 14C and 14D , the LVDT arrangement described above is modified by removing the second coil  898  in ( FIG. 14A ). When coil  898  is removed, the voltage induced in coil  897  will vary relative to the end position  874  of the core  873 . In embodiments where the direction and timing of movement of the valve  872  is known, only one secondary coil  897  is necessary to measure magnitude of movement. As noted above, the core  873  portion of the valve can be located and fabricated using several methods. For example, a weld at the end position  874  can join nickel base non-core material and iron base core material, a physical reduction in diameter can be used to locate end position  874  to vary magnetic flux in a specific location, or a slug of iron-based material can be inserted and located at the end position  874 . 
     It will be appreciated in light of the disclosure that the LVDT sensor components in one example can be located near the top of the valve guide  871  to allow for temperature dissipation below that point. While such a location can be above typical weld points used in valve stem fabrication, the weld could be moved or as noted. The location of the core  873  relative to the secondary winding  897  is proportional to how much voltage is induced. 
     The use of an LVDT sensor as described above in an operating engine has several advantages, including 1) Frictionless operation—in normal use, there is no mechanical contact between the LVDT&#39;s core  873  and coil assembly. No friction also results in long mechanical life. 2) Nearly infinite resolution—since an LVDT operates on electromagnetic coupling principles in a friction-free structure, it can measure infinitesimally small changes in core position, limited only by the noise in an LVDT signal conditioner and the output display&#39;s resolution. This characteristic also leads to outstanding repeatability, 3) Environmental robustness—materials and construction techniques used in assembling an LVDT result in a rugged, durable sensor that is robust to a variety of environmental conditions. Bonding of the windings  897 ,  898 ,  899  may be followed by epoxy encapsulation into the valve guide body  871 , resulting in superior moisture and humidity resistance, as well as the capability to take substantial shock loads and high vibration levels. Additionally, the coil assembly can be hermetically sealed to resist oil and corrosive environments. 4) Null point repeatability—the location of an LVDT&#39;s null point, described previously, is very stable and repeatable, even over its very wide operating temperature range. 5) Fast dynamic response—the absence of friction during ordinary operation permits an LVDT to respond very quickly to changes in core position. The dynamic response of an LVDT sensor is limited only by small inertial effects due to the core assembly mass. In most cases, the response of an LVDT sensing system is determined by characteristics of the signal conditioner. 6) Absolute output—an LVDT is an absolute output device, as opposed to an incremental output device. This means that in the event of loss of power, the position data being sent from the LVDT will not be lost. When the measuring system is restarted, the LVDT&#39;s output value will be the same as it was before the power failure occurred. 
     The valve stem position sensor described above employs a LVDT type transducer to determine the location of the valve stem during operation of the engine. The sensor may be any known sensor technology including Hall-effect sensor, electronic, optical and mechanical sensors that can track the position of the valve stem and report the monitored position back to the ECU. 
     2.4.3 Part Position/Movement 
     Variable valve actuation (VVA) technologies are designed to change valve lift profiles during engine operation using switching devices, for example a DVVL switching rocker arm. Changes in switching state may also change the position of component parts in VVA assemblies, either in absolute terms or relative to one another in the assembly. Position change measurements can be designed and implemented to monitor the state of VVA switching, and possibly determine if there is a switching malfunction. 
     Now, with reference to  FIGS. 15-16 , an exemplary DVVL switching rocker arm assembly  100  can be configured with an accurate non-contacting sensor  828  that measures relative movement, motion, or distance. 
     In one embodiment, movement sensor  828  is located near the first end  101  ( FIG. 15 ), to evaluate the movement of the outer arm  120  relative to known positions for high-lift and low-lift modes. In this example, movement sensor  828  comprises a wire wound around a permanently magnetized core, and is located and oriented to detect movement by measuring changes in magnetic flux produced as a ferrous material passes through its known magnetic field. For example, when the outer arm tie bar  875 , which is magnetic (ferrous material), passes through the permanent magnetic field of the position sensor  828 , the flux density is modulated, inducing AC voltages in the coil and producing an electrical output that is proportional to the proximity of the tie bar  875 . The modulating voltage is input to the engine control unit (ECU) (described in following sections), where a processor employs logic and calculations to initiate rocker arm assembly  100  switching operations. In embodiments, the voltage output may be binary, meaning that the absence or presence of a voltage signal indicates high-lift or low-lift. 
     It can be seen that position sensor  828  may be positioned to measure movement of other parts in the rocker arm assembly  100 . In a second embodiment, sensor  828  may be positioned at second end  103  of the DVVL rocker arm assembly  100  ( FIG. 15 ) to evaluate the location of the inner arm  122  relative to the outer arm  120 . 
     A third embodiment can position sensor  828  to directly evaluate the latch  200  position in the DVVL rocker arm assembly  100 . The latch  200  and sensor  828  are engaged and fixed relative to each other when they are in the latched state (high lift mode), and move apart for unlatched (low-lift) operation. 
     Movement may also be detected using and inductive sensor. Sensor  877  may be a Hall-effect sensor, mounted in a way that allows measurement of the movement or lack of movement, for example the valve  872  stem. 
     2.4.4 Pressure Characterization 
     Variable valve actuation (VVA) technologies are designed to change valve lift profiles during engine operation using switching devices, for example a DVVL switching rocker arm. Because latch status is an important input to the ECU that may enable it to perform various functions, such as regulating fuel/air mixture to increase gas mileage, reduce pollution, or to regulate idle and knocking, measuring devices or systems that confirm a successful switching operation, or detect an error condition or malfunction are necessary for proper control. In some cases switching status reporting and error notification is necessary for regulatory compliance. 
     In embodiments comprising a hydraulically actuated DVVL system  800 , as illustrated in  FIG. 6 , changes in switching state provide distinct hydraulic switching fluid pressure signatures. Because fluid pressure is required to produce the necessary hydraulic stiffness that initiates switching, and because hydraulic fluid pathways are geometrically defined with specific channels and chambers, a characteristic pressure signature is produced that can be used to predictably determine latched or unlatched status or a switching malfunction. Several embodiments can be described that measure pressure, and compare measured results with known and acceptable operating parameters. Pressure measurements can be analyzed on a macro level by examining fluid pressure over several switching cycles, or evaluated over a single switching event lasting milliseconds. 
     Now, with reference to  FIGS. 6, 7, and 17 , an example plot ( FIG. 17 ) shows the valve lift height variation  882  over time for cylinder  1  as the switching rocker assembly  100  operates in either high-lift or low-lift, and switches between high-lift and low-lift. Corresponding data for the hydraulic switching system are plotted on the same time scale ( FIG. 17 ), including oil pressure  880  in the upper galleries  802 ,  803  as measured using pressure transducer  890 , and the electrical current  881  used to open and close solenoid valves  822 ,  823  in the OCV assembly  820 . As can be seen, this level of analysis on a macro level clearly shows the correlation between OCV switching current  881 , control pressure  880 , and lift  882  during all states of operation. For example, at time 0.1, the OCV is commanded to switch, as shown by an increased electrical current  881 . When the OCV is switched, increased control pressure  880  results in a high-lift to low-lift switching event. As operation is evaluated over one or more complete switching cycles, proper operation of the sub-system comprising the OCV and the pressurized fluid delivery system to the rocker arm assembly  100  can be evaluated. Switching malfunction determination can be enhanced with other independent measurements, for example valve stem movement as described above. As can be seen, these analyses can be performed for any number of OCV&#39;s used to control intake and/or exhaust valves for one or more cylinders. 
     Using a similar method, but using data measured and analyzed on the microsecond level during a switching event, provides enough detailed control pressure information ( FIGS. 17A, 17B ) to independently evaluate a successful switching event or switching malfunction without measuring valve lift or latch pin movement directly. In embodiments using this method, switching state is determined by comparing the measured pressure transient to known operating state pressure transients developed during testing, and stored in the ECU for analysis.  FIGS. 17A and 17B  illustrate exemplary test data used to produce known operating pressure transients for a switching rocker arm in a DVVL system. 
     The test system included four switching rocker arm assemblies  100  as shown in ( FIG. 3 ), an OCV assembly  820  ( FIG. 3 ), two upper oil control galleries  802 ,  803  ( FIGS. 6-7 ), and a closed loop system to control hydraulic actuating fluid temperature and pressure in the control galleries  802 ,  803 . Each control gallery provided hydraulic fluid at regulated pressure to control two rocker arm assemblies  100 .  FIG. 17A  illustrates a valid single test run showing data when an OCV solenoid valve is energized to initiate switching from high-lift to low-lift state. Instrumentation was installed to measure latch movements  1002 , pressure  880  in the control galleries  802 ,  803 , OCV current  881 , pressure  1001  in the hydraulic fluid supply  804  ( FIG. 6-7 ), and latch lash and cam lash. The sequence of events can be described as follows:
         0 ms—ECU switched on electrical current  881  to energize the OCV solenoid valve.   10 ms—Switching current  881  to the OCV solenoid is sufficient to regulate pressure higher in the control gallery  802 ,  803  as shown by pressure curve  880 .   10-13 ms—The supply pressure curve  1001  decreases below the pressure regulated by the OCV as hydraulic fluid flows from the supply  804  ( FIGS. 6-7 ) into the upper control galleries  802 ,  803 . In response, pressure  880  increases rapidly in the control galleries  802 ,  803 . Latch pin movement begins as shown in latch pin movement curve  1002 .   13-15 ms—The supply pressure curve  1001  returns to a steady unregulated state as flow stabilizes. Pressure  880  in the control galleries  802 ,  803  increases to the higher pressure regulated by the OCV.   15-20 ms—A pressure  880  increase/decrease transient in the control galleries  802 ,  803  is produced as pressurized hydraulic fluid pushes the latch fully back into position (latch pin movement curve  1002 ), and hydraulic flow and pressure stabilizes at the OCV unregulated pressure. Pressure spike  1003  is characteristic of this transient.   At 12 ms and 17 ms distinctive pressure transients can be seen in pressure curve  880  that coincide with sudden changes in latch position  1002 .       

       FIG. 17B  illustrates a valid single test run showing data when an OCV solenoid valve is de-energized to initiate switching from low-lift to high-lift state. The sequence of events can be described as follows:
         0 ms—ECU switched off electrical current  881  to de-energize the OCV solenoid valve.   5 ms—OCV solenoid moves far enough to introduce regulated, lower pressure, hydraulic fluid into enter the control galleries  802  and  803  (pressure curve  880 ).   5-7 ms—Pressure in the control galleries  802 ,  803 , decreases rapidly as shown by curve  880 , as the OCV regulates pressure lower.   7-12 ms—Coinciding with the low pressure point  1005 , lower pressure in the control galleries  802 ,  803  initiates latch movement as shown by the latch movement curve  1002 . Pressure curve  880  transients are initiated as the latch spring  230  ( FIG. 19 ) compresses and moves hydraulic fluid in the volume engaging the latch.   12-15 ms—Pressure transients, shown in pressure curve  880 , are again introduced as the latch pin movement, shown by latch pin movement curve  1002 , completes.   15-30 ms—Pressure in control galleries  802 ,  803  stabilize at the OCV regulated pressure as shown by pressure curve  880 .   As noted above, at 7-10 ms and 13-20 ms distinctive pressure transients can be seen in pressure curve  880  that coincide with sudden changes in latch position  1002 .       

     As noted previously, and in following sections, the fixed geometric configuration of the hydraulic channels, holes, clearances, and chambers, and the stiffness of the latch spring, are variables that relate to hydraulic response and mechanical switching speed for changes in regulated hydraulic fluid pressure. The pressure curves  880 , in  FIGS. 17A and 17B  describe a DVVL switching rocker arm system operating in an acceptable range. During operation, specific rates of increase or decrease in pressure (curve slope) are characteristic of proper operation characterized by the timing of events listed above. Examples of error conditions include: time shifting of pressure events that show deterioration of latch response times, changes in rate of the occurrence of events (pressure curve slope changes), or an overall decrease in the amplitude of pressure events. For example, a lower than anticipated pressure increase in the 15-20 ms period indicates that the latch has not retracted completely, potentially resulting in a critical shift. 
     The test data in these examples were measured with oil pressure of 50 psi and oil temperature of 70 degrees C. A series of tests in different operating conditions can provide a database of characteristic curves to be used by the ECU for switching diagnosis. 
     An additional embodiment that utilizes pressure measurement to diagnose switching state is described. A DFHLA  110  as shown in  FIG. 3  is used to both manage lash, and supply hydraulic fluid for actuating VVA systems that employ switching rocker arm assemblies such as CDA or DVVL. As shown in the section view of  FIG. 52 , normal lash adjustment for the DVVL rocker arm assembly  100 , causes the ball plunger  601  to maintain contact with the receiving socket of the inner arm assembly  622  during both high-lift and low-lift operation. When fully assembled in an engine, the DFHLA  110  is in a fixed position, while the inner rocker arm assembly  622  exhibits rotational movement about the ball tip contact point  611 . The rotational movement of the inner arm assembly  622  and the ball plunger load  615  vary in magnitude when switching between high-lift and low-lift states. The ball plunger  601  is designed to move in compensation when loads and movement vary. 
     Compensating force for the ball plunger load  615  is provided by hydraulic fluid pressure in the lower control gallery  805  as it is communicated from the lower port  512  to chamber  905  ( FIG. 11 ). As shown in  FIGS. 6-7 , hydraulic fluid at unregulated pressure is communicated from the engine cylinder head, into the lower control gallery  805 . 
     In embodiments, a pressure transducer is placed in the hydraulic gallery  805  that feeds the lash adjuster part of the DFHLA  110 . The pressure transducer can be used to monitor the transient pressure change in the hydraulic gallery  805  that feeds the lash adjuster when transitioning from the high-lift state to the low-lift state or from the low-lift state to the high-lift state. By monitoring the pressure signature when switching from one mode to another, the system may be able to detect when the variable valve actuation system is malfunctioning at any one location. A pressure signature curve, in embodiments plotted as pressure versus time in milliseconds, provides a characteristic shape that can include amplitude, slope, and/or other parameters. 
     For example,  FIG. 17C  shows a plot of intake valve lift profile curves  814 ,  816  versus time in milliseconds, superimposed with a plot of hydraulic gallery pressure curves  1005 ,  1006  versus the same time scale. Pressure curve  1006  and valve lift profile curve  816  correspond to the low-lift state, and pressure curve  1005  and valve lift profile  814  correspond to the high-lift state. 
     During steady state operation, pressure signature curves  1005 ,  1006  exhibit cyclical behavior, with distinct spikes  1007 ,  1008  caused as the DFHLA compensates for alternating ball plunger loads  615  that are imparted as the cam pushes down the rocker arm assembly to compress the valve spring ( FIG. 3 ) and provide valve lift, as the valve spring extends to close the valve, and when the cam is on base circle where no lift occurs. As shown in  FIG. 17C , transient pressure spikes  1008 ,  1007  correspond with the peak of the low-lift and high-lift profiles  816 ,  814  respectively. As the hydraulic system pressure stabilizes, steady-state pressure signature curves  1005 ,  1006  resume. 
     As noted previously, and in following sections, the fixed geometric configuration of DFHLA hydraulic channels, holes, clearances, and chambers, are variables that relate to hydraulic response and pressure transients for a given hydraulic fluid pressure and temperature. The pressure signature curves  1005 ,  1006 , in  FIG. 17C  describe a DVVL switching rocker arm system operating in an acceptable range. During operation, certain rates of increase or decrease in pressure (curve slopes), peak pressure values, and timing of peak pressures with respect to maximum lift are also be characteristic of proper operation characterized by the timing of switching events. Examples of error conditions may include time shifting of pressure events, changes in rate of the occurrence of events (pressure curve slope changes), sudden unexpected pressure transients, or an overall decrease in the amplitude of pressure events. 
     A series of tests in different operating conditions can provide a database of characteristic curves to be used by the ECU for switching diagnosis. One or several values of pressure can be used based on the system configuration and vehicle demands. The monitored pressure trace can be compared to a standard trace to determine when the system malfunctions. 
     3. Switching Control and Logic 
     3.1. Engine Implementation 
     The DVVL hydraulic fluid system that delivers engine oil at a controlled pressure to the DVVL switching rocker arm  100 , illustrated in  FIG. 4 , is described in following sections as it may be installed on an intake valve in a Type II valve train in a four cylinder engine. In alternate embodiments, this hydraulic fluid delivery system can be applied to any combination of intake or exhaust valves on a piston-driven internal combustion engines. 
     3.2. Hydraulic Fluid Delivery System to the Rocker Arm Assembly 
     With reference to  FIGS. 3, 6 and 7 , the hydraulic fluid system delivers engine oil at a controlled pressure to the DVVL switching rocker arm  100  ( FIG. 4 ). In this arrangement, engine oil from the cylinder head  801  that is not pressure regulated feeds into the HLA lower feed gallery  805 . As shown in  FIG. 3 , this oil is always in fluid communication with the lower feed inlet  512  of the DFHLA, where it is used to perform normal hydraulic lash adjustment. Engine oil from the cylinder head  801  that is not pressure regulated is also supplied to the oil control valve assembly inlet  821 . As described previously, the OCV assembly  820  for this DVVL embodiment comprises two independently actuated solenoid valves that regulate oil pressure from the common inlet  821 . Hydraulic fluid from the OCV assembly  820  first control port outlet  822  is supplied to the first upper gallery  802 , and hydraulic fluid from the second control port  823  is supplied to the second upper gallery  803 . The first OCV determines the lift mode for cylinders one and two, and the second OCV determines the lift mode for cylinders three and four. As shown in  FIG. 18  and described in following sections, actuation of valves in the OCV assembly  820  is directed by the engine control unit  825  using logic based on both sensed and stored information for particular physical configuration, switching window, and set of operating conditions, for example, a certain number of cylinders and a certain oil temperature. Pressure regulated hydraulic fluid from the upper galleries  802 ,  803  is directed to the DFHLA upper port  506 , where it is transmitted through channel  509  to the switching rocker arm assembly  100 . As shown in  FIG. 19 , hydraulic fluid is communicated through the rocker arm assembly  100  via the first oil gallery  144 , and the second oil gallery  146  to the latch pin assembly  201 , where it is used to initiate switching between high-lift and low-lift states. 
     Purging accumulated air in the upper galleries  802 ,  803  is important to maintain hydraulic stiffness and minimize variation in the pressure rise time. Pressure rise time directly affects the latch movement time during switching operations. The passive air bleed ports  832 ,  833  shown in  FIG. 6  were added to the high points in the upper galleries  802 ,  803  to vent accumulated air into the cylinder head air space under the valve cover. 
     3.2.1 Hydraulic Fluid Delivery for Low-Lift Mode: 
     Now, with reference to  FIG. 8 , the DVVL system is designed to operate from idle to 3500 rpm in low-lift mode. A section view of the rocker arm assembly  100  and the 3-lobed cam  102  shows low-lift operation. Major components of the assembly shown in  FIGS. 8 and 19 , include the inner arm  122 , roller bearing  128 , outer arm  120 , slider pads  130 ,  132 , latch  200 , latch spring  230 , pivot axle  118 , and lost motion torsion springs  134 ,  136 . For low-lift operation, when a solenoid valve in the OCV assembly  820  is energized, unregulated oil pressure at ≥2.0 Bar is supplied to the switching rocker arm assembly  100  through the control galleries  802 ,  803  and the DFHLA  110 . The pressure causes the latch  200  to retract, unlocking the inner arm  122  and outer arm  120 , and allowing them to move independently. The high-lift camshaft lobes  104 ,  106  ( FIG. 3 ) remain in contact with the sliding interface pads  130 ,  132  on the outer arm  120 . The outer arm  120  pivots about the pivot axle  118  and does not impart any motion to the valve  112 . This is commonly referred to as lost motion. Since the low-lift cam profile  816  ( FIG. 5 ) is designed for early valve closing, the switching rocker arm  100  must be designed to absorb all of the motion from the high-lift camshaft lobes  104 ,  106  ( FIG. 3 ). Force from the lost motion torsion springs  134 ,  136  ( FIG. 15 ) ensure the outer arm  120  stays in contact with the high-lift lobe  104 ,  106  ( FIG. 3 ). The low-lift lobe  108  ( FIG. 3 ) contacts the roller bearing  128  on the inner arm  122  and the valve is opened per the low lift early valve closing profile  816  ( FIG. 5 ). 
     3.2.2 Hydraulic Fluid Delivery for High-Lift Mode 
     Now, with reference to  FIG. 9 , the DVVL system is designed to operate from idle to 7300 rpm in high-lift mode. A section view of the switching rocker arm  100  and the 3-lobe cam  102  shows high-lift operation. Major components of the assembly are shown in  FIGS. 9 and 19 , including the inner arm  122 , roller bearing  128 , outer arm  120 , slider pads  130 ,  132 , latch  200 , latch spring  230 , pivot axle  118 , and lost motion torsion springs  134 ,  136 . 
     Solenoid valves in the OCV assembly  820  are de-energized to enable high lift operation. The latch spring  230  extends the latch  200 , locking the inner arm  122  and outer arm  120 . The locked arms function like a fixed rocker arm. The symmetric high lift lobes  104 ,  106  ( FIG. 3 ) contact the slider pads  130 , ( 132  not shown) on the outer arm  120 , rotating the inner arm  122  about the DFHLA  110  ball end  601  and opening the valve  112  ( FIG. 4 ) per the high lift profile  814  ( FIG. 5 ). During this time, regulated oil pressure from 0.2 to 0.4 bar is supplied to the switching rocker arm  100  through the control galleries  802 ,  803 . Oil pressure maintained at 0.2 to 0.4 bar keeps the oil passages full but does not retract the latch  200 . 
     In high-lift mode, the dual feed function of the DFHLA is important to ensure proper lash compensation of the valve train at maximum engine speeds. The lower gallery  805  in  FIG. 9  communicates cylinder head oil pressure to the lower DFHLA port  512  ( FIG. 11 ). The lower portion of the DFHLA is designed to perform as a normal hydraulic lash compensation mechanism. The DFHLA  110  mechanism was designed to ensure the hydraulics have sufficient pressure to avoid aeration and to remain full of oil at all engine speeds. Hydraulic stiffness and proper valve train function are maintained with this system. 
     The table in  FIG. 20  summarizes the pressure states in high-lift and low-lift modes. Hydraulic separation of the DFHLA normal lash compensation function from the rocker arm assembly switching function is also shown. The engine starts in high-lift mode (latch extended and engaged), since this is the default mode. 
     3.3 Operating Parameters 
     An important factor in operating a DVVL system is the reliable control of switching from high-lift mode to low-lift mode. DVVL valve actuation systems can only be switched between modes during a predetermined window of time. As described above, switching from high lift mode to low lift mode and vice versa is initiated by a signal from the engine control unit (ECU)  825  ( FIG. 18 ) using logic that analyzes stored information, for example a switching window for particular physical configuration, stored operating conditions, and processed data that is gathered by sensors. Switching window durations are determined by the DVVL system physical configuration, including the number of cylinders, the number of cylinders controlled by a single OCV, the valve lift duration, engine speed, and the latch response times inherent in the hydraulic control and mechanical system. 
     3.3.1 Gathered Data 
     Real-time sensor information includes input from any number of sensors, as illustrated in the exemplary DVVL system  800  illustrated in  FIG. 6 . Sensors may include 1) valve stem movement  829 , as measured in one embodiment using the linear variable differential transformer (LVDT) described previously, 2) motion/position  828  and latch position  827  using a Hall-effect sensor or motion detector, 3) DFHLA movement  826  using a proximity switch, Hall effect sensor, or other means, 4) oil pressure  830 , and 5) oil temperature  890 . Cam shaft rotary position and speed may be gathered directly or inferred from the engine speed sensor. 
     In a hydraulically actuated VVA system, the oil temperature affects the stiffness of the hydraulic system used for switching in systems such as CDA and VVL. If the oil is too cold, its viscosity slows switching time, causing a malfunction. This relationship is illustrated for an exemplary DVVL switching rocker arm system, in  FIGS. 21-22 . An accurate oil temperature, taken with a sensor  890  shown in  FIG. 6 , located near the point of use rather than in the engine oil crankcase, provides the most accurate information. In one example, the oil temperature in a VVA system, monitored close to the oil control valves (OCV), must be greater than or equal to 20 degrees C. to initiate low-lift (unlatched) operation with the required hydraulic stiffness. Measurements can be taken with any number of commercially available components, for example a thermocouple. The oil control valves are described further in published US Patent Applications US2010/0089347 published Apr. 15, 2010 and US2010/0018482 published Jan. 28, 2010 both hereby incorporated by reference in their entirety. 
     Sensor information is sent to the Engine Control Unit (ECU)  825  as a real-time operating parameter ( FIG. 18 ). 
     3.3.2 Stored Information 
     3.3.2.1 Switching Window Algorithms 
     Mechanical Switching Window: 
     The shape of each lobe of the three-lobed cam illustrated in  FIG. 4  comprises a base circle portion  605 ,  607 ,  609 , where no lift occurs, a transition portion that is used to take up mechanical clearances prior to a lift event, and a lift portion that moves the valve  112 . For the exemplary DVVL switching rocker arm  100 , installed in system  800  ( FIG. 6 ), switching between high-lift and low-lift modes can only occur during base circle operation when there is no load on the latch that prevents it from moving. Further descriptions of this mechanism are provided in following sections. The no-lift portion  863  of base circle operation is shown graphically in  FIG. 5 . The DVVL system  800 , switches within a single camshaft revolution at speeds up to 3500 engine rpm at oil temperatures of 20° C. and above. Switching outside of the timing window or prescribed oil conditions may result in a critical shift event, which is a shift in engine valve position during a point in the engine cycle when loading on the valve actuator switching component or on the engine valve is higher than the structure is designed to accommodate while switching. A critical shift event may result in damage to the valve train and/or other engine parts. The switching window can be further defined as the duration in cam shaft crank degrees needed to change the pressure in the control gallery and move the latch from the extended to retracted position and vice versa. 
     As previously described and shown in  FIG. 7 , the DVVL system has a single OCV assembly  820  that contains two independently controlled solenoid valves. The first valve controls the first upper gallery  802  pressure and determines the lift mode for cylinders one and two. The second valve controls the second upper gallery  803  pressure and determines the lift mode for cylinders three and four.  FIG. 23  illustrates the intake valve timing (lift sequence) for this OCV assembly  820  ( FIG. 3 ) configuration relative to crankshaft angle for an in-line four cylinder engine with a cylinder firing order of (2-1-3-4). The high-lift intake valve profiles for cylinder two  851 , cylinder one  852 , cylinder three  853 , and cylinder four  854 , are shown at the top of the illustration as lift plotted versus crank angle. Valve lift duration for the corresponding cylinders are plotted in the lower section as lift duration regions  855 ,  856 ,  857 , and  858  lift versus crank angle. No lift base circle operating regions  863  for individual cylinders are also shown. A prescribed switching window must be determined to move the latch within one camshaft revolution, with the stipulation that each OCV is configured to control two cylinders at once. 
     The mechanical switching window can be optimized by understanding and improving latch movement. Now, with reference to  FIGS. 24-25 , the mechanical configuration of the switching rocker arm assembly  100  provides two distinct conditions that allow the effective switching window to be increased. The first, called a high-lift latch restriction, occurs in high-lift mode when the latch  200  is locked in place by the load being applied to open the valve  112 . The second, called a low-lift latch restriction, occurs in the unlatched low-lift mode when the outer arm  120  blocks the latch  200  from extending under the outer arm  120 . These conditions are described as follows: 
     High-Lift Latch Restriction: 
       FIG. 24  shows high-lift event where the latch  200  is engaged with the outer arm  120 . As the valve is opened against the force supplied by valve spring  114 , the latch  200  transfers the force from the inner arm  122  to the outer arm  120 . When the spring  114  force is transferred by the latch  200 , the latch  200  becomes locked in its extended position. In this condition, hydraulic pressure applied by switching the OCV while attempting to switch from high-lift to low-lift mode is insufficient to overcome the force locking the latch  200 , preventing it from being retracted. This condition extends the total switching window by allowing pressure application prior to the end of the high-lift event and the onset of base circle  863  ( FIG. 23 ) operation that unloads the latch  200 . When the force is released on the latch  200 , a switching event can commence immediately. 
     Low-Lift Latch Restriction: 
       FIG. 25  shows low lift operation where the latch  200  is retracted in low-lift mode. During the lift portion of the event, the outer arm  120  blocks the latch  200 , preventing its extension, even if the OCV is switched, and hydraulic fluid pressure is lowered to return to the high-lift latched state. This condition extends the total switching window by allowing hydraulic pressure release prior to the end of the high-lift event and the onset of base circle  863  ( FIG. 23 ). Once base circle is reached, the latch spring  230  can extend the latch  200 . The total switching window is increased by allowing pressure relief prior to base circle. When the camshaft rotates to base circle, switching can commence immediately. 
       FIG. 26  illustrates the same information shown in  FIG. 23 , but is also overlaid with the time required to complete each step of the mechanical switching process during the transition between high-lift and low-lift states. These steps represent elements of mechanical switching that are inherent in the design of the switching rocker arm assembly. As described for  FIG. 23 , the firing order of the engine is shown at the top corresponding to the crank angle degrees referenced to cylinder two along with the intake valve profiles  851 ,  852 ,  853 ,  854 . The latch  200  must be moved while the intake cam lobes are on base circle  863  (referred to as the mechanical switching window). Since each solenoid valve in an OCV assembly  820  controls two cylinders, the switching window must be timed to accommodate both cylinders while on their respective base circles. Cylinder two returns to base circle at 285 degrees crank angle. Latch movement must be complete by 690 crank angle degrees prior to the next lift event for cylinder two. Similarly, cylinder one returns to base circle at 465 degrees and must complete switching by 150 degrees. As can be seen, the switching window for cylinders one and two is slightly different. As can be seen, the first OCV electrical trigger starts switching prior to the cylinder one intake lift event and the second OCV electrical trigger starts prior to the cylinder four intake lift event. 
     A worst case analysis was performed to define the switching times in  FIG. 26  at the maximum switching speed of 3500 rpm. Note that the engine may operate at much higher speeds of 7300 rpm; however, mode switching is not allowed above 3500 rpm. The total switching window for cylinder two is 26 milliseconds, and is broken into two parts: a 7 millisecond high-lift/low-lift latch restriction time  861 , and a 19 millisecond mechanical switching time  864 . A 10 millisecond mechanical response time  862  is consistent for all cylinders. The 15 millisecond latch restricted time  861  is longer for cylinder one because OCV switching is initiated while cylinder one is on an intake lift event, and the latch is restricted from moving. 
     Several mechanical and hydraulic constraints that must be accommodated to meet the total switching window. First, a critical shift  860 , caused by switching that is not complete prior to the beginning of the next intake lift event must be avoided. Second, experimental data shows that the maximum switching time to move the latch at the lowest allowable engine oil temperature of 20° C. is 10 milliseconds. As noted in  FIG. 26 , there are 19 milliseconds available for mechanical switching  864  on the base circle. Because all test data shows that the switching mechanical response  862  will occur in the first 10 milliseconds, the full 19 milliseconds of mechanical switching time  864  is not required. The combination of mechanical and hydraulic constraints defines a worst-case switching time of 17 milliseconds that includes latch restricted time  861  plus latch mechanical response time  862 . 
     The DVVL switching rocker arm system was designed with margin to accomplish switching with a 9 millisecond margin. Further, the 9 millisecond margin may allow mode switching at speeds above 3500 rpm. Cylinders three and four correspond to the same switching times as one and two with different phasing as shown in  FIG. 26 . Electrical switching time required to activate the solenoid valves in the OCV assembly is not accounted for in this analysis, although the ECU can easily be calibrated to consider this variable because the time from energizing the OCV until control gallery oil pressure begins to change remains predictable. 
     Now, as to  FIGS. 4 and 25A , a critical shift may occur if the timing of the cam shaft rotation and the latch  200  movement coincide to load the latch  200  on one edge, where it only partially engages on the outer arm  120 . Once the high-lift event begins, the latch  200  can slip and disengage from the outer arm  120 . When this occurs, the inner arm  122 , accelerated by valve spring  114  forces, causes an impact between the roller bearing  128  and the low-lift cam lobe  108 . A critical shift is not desired as it creates a momentary loss of control of the rocker arm assembly  100  and valve movement, and an impact to the system. The DVVL switching rocker arm was designed to meet a lifetime worth of critical shift occurrences. 
     3.3.2.2 Stored Operating Parameters 
     Operating parameters comprise stored information, used by the ECU  825  ( FIG. 18 ) for switching logic control, based on data collected during extended testing as described in later sections. Several examples of known operating parameters may be described: In embodiments, 1) a minimum oil temperature of 20 degrees C. is required for switching from a high-lift state to a low-lift state, 2) a minimum oil pressure of greater than 2 Bar should be present in the engine sump for switching operations, 3) The latch response switching time varies with oil temperature according to data plotted in  FIGS. 21-22 , 4) as shown in  FIG. 17  and previously described, predictable pressure variations caused by hydraulic switching operations occur in the upper galleries  802 ,  803  ( FIG. 6 ) as determined by pressure sensors  890 , 5) as shown in  FIG. 5  and previously described, known valve movement versus crank angle (time), based on lift profiles  814 ,  816  can be predetermined and stored. 
     3.4 Control Logic 
     As noted above, DVVL switching can only occur during a small predetermined window of time under certain operating conditions, and switching the DVVL system outside of the timing window may result in a critical shift event, that could result in damage to the valve train and/or other engine parts. Because engine conditions such as oil pressure, temperature, emissions, and load may vary rapidly, a high-speed processor can be used to analyze real-time conditions, compare them to known operating parameters that characterize a working system, reconcile the results to determine when to switch, and send a switching signal. These operations can be performed hundreds or thousands of times per second. In embodiments, this computing function may be performed by a dedicated processor, or by an existing multi-purpose automotive control system referred to as the engine control unit (ECU). A typical ECU has an input section for analog and digital data, a processing section that includes a microprocessor, programmable memory, and random access memory, and an output section that might include relays, switches, and warning light actuation. 
     In one embodiment, the engine control unit (ECU)  825  shown in  FIGS. 6 and 18  accepts input from multiple sensors such as valve stem movement  829 , motion/position  828 , latch position  827 , DFHLA movement  826 , oil pressure  830 , and oil temperature  890 . Data such as allowable operating temperature and pressure for given engine speeds ( FIG. 20 ), and switching windows ( FIG. 26  and described in other sections), is stored in memory. Real-time gathered information is then compared with stored information and analyzed to provide the logic for ECU  825  switching timing and control. 
     After input is analyzed, a control signal is output by the ECU  825  to the OCV  820  to initiate switching operation, which may be timed to avoid critical shift events while meeting engine performance goals such as improved fuel economy and lowered emissions. If necessary, the ECU  825  may also alert operators to error conditions. 
     4. DVVL Switching Rocker Arm Assembly 
     4.1 Assembly Description 
     A switching rocker arm, hydraulically actuated by pressurized fluid, for engaging a cam is disclosed. An outer arm and inner arm are configured to transfer motion to a valve of an internal combustion engine. A latching mechanism includes a latch, sleeve and orientation member. The sleeve engages the latch and a bore in the inner arm, and also provides an opening for an orientation member used in providing the correct orientation for the latch with respect to the sleeve and the inner arm. The sleeve, latch and inner arm have reference marks used to determine the optimal orientation for the latch. 
     An exemplary switching rocker arm  100  may be configured during operation with a three lobed cam  102  as illustrated in the perspective view of  FIG. 4 . Alternatively, a similar rocker arm embodiment could be configured to work with other cam designs such as a two lobed cam. The switching rocker arm  100  is configured with a mechanism to maintain hydraulic lash adjustment and a mechanism to feed hydraulic switching fluid to the inner arm  122 . In embodiments, a dual feed hydraulic lash adjuster (DFHLA)  110  performs both functions. A valve  112 , spring  114 , and spring retainer  116  are also configured with the assembly. The cam  102  has a first and second high-lift lobe  104 ,  106  and a low lift lobe  108 . The switching rocker arm has an outer arm  120  and an inner arm  122 , as shown in  FIG. 27 . During operation, the high-lift lobes  104 ,  106  contact the outer arm  120  while the low lift-lobe contacts the inner arm  122 . The lobes cause periodic downward movement of the outer arm  120  and inner arm  122 . The downward motion is transferred to the valve  112  by inner arm  122 , thereby opening the valve. Rocker arm  100  is switchable between a high-lift mode and low-lift mode. In the high-lift mode, the outer arm  120  is latched to the inner arm  122 . During engine operation, the high-lift lobes periodically push the outer arm  120  downward. Because the outer arm  120  is latched to the inner arm  122 , the high-lift motion is transferred from outer arm  120  to inner arm  122  and further to the valve  112 . When the rocker arm  100  is in its low-lift mode, the outer arm  120  is not latched to the inner arm  122 , and so high-lift movement exhibited by the outer arm  120  is not transferred to the inner arm  122 . Instead, the low-lift lobe contacts the inner arm  122  and generates low lift motion that is transferred to the valve  112 . When unlatched from inner arm  122 , the outer arm  120  pivots about axle  118 , but does not transfer motion to valve  112 . 
       FIG. 27  illustrates a perspective view of an exemplary switching rocker arm  100 . The switching rocker arm  100  is shown by way of example only and it will be appreciated that the configuration of the switching rocker arm  100  that is the subject of this disclosure is not limited to the configuration of the switching rocker arm  100  illustrated in the figures contained herein. 
     As shown in  FIG. 27 , the switching rocker arm  100  includes an outer arm  120  having a first outer side arm  124  and a second outer side arm  126 . An inner arm  122  is disposed between the first outer side arm  124  and second outer side arm  126 . The inner arm  122  and outer arm  120  are both mounted to a pivot axle  118 , located adjacent the first end  101  of the rocker arm  100 , which secures the inner arm  122  to the outer arm  120  while also allowing a rotational degree of freedom about the pivot axle  118  of the inner arm  122  with respect to the outer arm  120 . In addition to the illustrated embodiment having a separate pivot axle  118  mounted to the outer arm  120  and inner arm  122 , the pivot axle  118  may be part of the outer arm  120  or the inner arm  122 . 
     The rocker arm  100  illustrated in  FIG. 27  has a roller bearing  128  that is configured to engage a central low-lift lobe of a three-lobed cam. First and second slider pads  130 ,  132  of outer arm  120  are configured to engage the first and second high-lift lobes  104 ,  106  shown in  FIG. 4 . First and second torsion springs  134 ,  136  function to bias the outer arm  120  upwardly after being displaced by the high-lift lobes  104 ,  106 . The rocker arm design provides spring over-torque features. 
     First and second over-travel limiters  140 ,  142  of the outer arm prevent over-coiling of the torsion springs  134 ,  136  and limit excess stress on the springs  134 ,  136 . The over-travel limiters  140 ,  142  contact the inner arm  122  on the first and second oil gallery  144 ,  146  when the outer arm  120  reaches its maximum rotation during low-lift mode. At this point, the interference between the over-travel limiters  140 ,  142  and the galleries  144 ,  146  stops any further downward rotation of the outer arm  120 .  FIG. 28  illustrates a top-down view of rocker arm  100 . As shown in  FIG. 28 , over-travel limiters  140 ,  142  extend from outer arm  120  toward inner arm  122  to overlap with galleries  144 ,  146  of the inner arm  122 , ensuring interference between limiters  140 ,  142  and galleries  144 ,  146 . As shown in  FIG. 29 , representing a cross-section view taken along line  29 - 29 , contacting surface  143  of limiter  140  is contoured to match the cross-sectional shape of gallery  144 . This assists in applying even distribution of force when limiters  140 ,  142  make contact with galleries  144 ,  146 . 
     When the outer arm  120  reaches its maximum rotation during low-lift mode as described above, a latch stop  90 , shown in  FIG. 15 , prevents the latch from extending, and locking incorrectly. This feature can be configured as necessary, suitable to the shape of the outer arm  120 . 
       FIG. 27  shows a perspective view from above of a rocker assembly  100  showing torsion springs  134 ,  136  according to one embodiment of the teachings of the present application.  FIG. 28  is a plan view of the rocker assembly  100  of  FIG. 27 . This design shows the rocker arm assembly  100  with torsion springs  134 ,  136  each coiled around a retaining axle  118 . 
     The switching rocker arm assembly  100  must be compact enough to fit in confined engine spaces without sacrificing performance or durability. Traditional torsion springs coiled from round wire sized to meet the torque requirements of the design, in some embodiments, are too wide to fit in the allowable spring space  121  between the outer arm  120  and the inner arm  122 , as illustrated in  FIG. 28 . 
     4.2 Torsion Spring 
     A torsion spring  134 ,  136  design and manufacturing process is described that results in a compact design with a generally rectangular shaped wire made with selected materials of construction. 
     Now, with reference to  FIGS. 15, 28, 30A, and 30B , the torsion springs  134 ,  136 , are constructed from a wire  397  that is generally trapezoidal in shape. The trapezoidal shape is designed to allow wire  397  to deform into a generally rectangular shape as force is applied during the winding process. After torsion spring  134 ,  136  is wound, the shape of the resulting wires can be described as similar to a first wire  396  with a generally rectangular shape cross section. A section along line  8  in  FIG. 28  shows two torsion spring  134 ,  136  embodiments, illustrated as multiple coils  398 ,  399  in cross section. In a preferred embodiment, wire  396  has a rectangular cross sectional shape, with two elongated sides, shown here as the vertical sides  402 ,  404  and a top  401  and bottom  403 . The ratio of the average length of side  402  and side  404  to the average length of top  401  and bottom  403  of the coil can be any value less than 1. This ratio produces more stiffness along the coil axis  419  of bending than a spring coiled with round wire with a diameter equal to the average length of top  401  and bottom  403  of the coil  398 . In an alternate embodiment, the cross section wire shape has a generally trapezoidal shape with a larger top  401  and a smaller bottom  403 . 
     In this configuration, as the coils are wound, elongated side  402  of each coil rests against the elongated side  402  of the previous coil, thereby stabilizing the torsion springs  134 ,  136 . The shape and arrangement holds all of the coils in an upright position, preventing them from passing over each other or angling when under pressure. 
     When the rocker arm assembly  100  is operating, the generally rectangular or trapezoidal shape of the torsion springs  134 ,  136 , as they bend about axis  419  shown in  FIGS. 30A, 30B , and  FIG. 19 , produces high part stress, particularly tensile stress on top surface  401 . 
     To meet durability requirements, a combination of techniques and materials are used together. For example, the torsion springs  134 ,  136  may be made of a material that includes Chrome Vanadium alloy steel along with this design to improve strength and durability. 
     The torsion spring  134 ,  136  may be heated and quickly cooled to temper the springs. This reduces residual part stress. 
     Impacting the surface of the wire  396 ,  397  used for creating the torsion springs  134 ,  136  with projectiles, or ‘shot peening’ is used to put residual compressive stress in the surface of the wire  396 ,  397 . The wire  396 ,  397  is then wound into the torsion springs  134 ,  136 . Due to their shot peening, the resulting torsion springs  134 ,  136  can now accept more tensile stress than identical springs made without shot peening. 
     4.3 Torsion Spring Pocket 
     The switching rocker arm assembly  100  may be compact enough to fit in confined engine spaces with minimal impact to surrounding structures. 
     A switching rocker arm  100  provides a torsion spring pocket with retention features formed by adjacent assembly components is described. 
     Now with reference to  FIGS. 27, 19, 28, and 31 , the assembly of the outer arm  120  and the inner arm  122  forms the spring pocket  119  as shown in  FIG. 31 . The pocket includes integral retaining features  119  for the ends of torsion springs  134 ,  136  of  FIG. 19 . 
     Torsion springs  134 ,  136  can freely move along the axis of pivot axle  118 . When fully assembled, the first and second tabs  405 ,  406  on inner arm  122  retain inner ends  409 ,  411  of torsion springs  134 ,  136 , respectively. The first and second over-travel limiters  140 ,  142  on the outer arm  120  assemble to prevent rotation and retain outer ends  407 ,  408  of the first and second torsion springs  134 ,  136 , respectively, without undue constraints or additional materials and parts. 
     4.4 Outer Arm 
     The design of outer arm  120  is optimized for the specific loading expected during operation, and its resistance to bending and torque applied by other means or from other directions may cause it to deflect out of specification. Examples of non-operational loads may be caused by handling or machining A clamping feature or surface built into the part, designed to assist in the clamping and holding process while grinding the slider pads, a critical step needed to maintain parallelism between the slider pads as it holds the part stationary without distortion.  FIG. 15  illustrates another perspective view of the rocker arm  100 . A first clamping lobe  150  protrudes from underneath the first slider pad  130 . A second clamping lobe (not shown) is similarly placed underneath the second slider pad  132 . During the manufacturing process, clamping lobes  150  of outer arm  120  are engaged by a clamping fixture  165  engaging outer arm  120  with clamps  175  during grinding of the slider pads  130 ,  132 . Forces are applied to the clamping lobes  150  by the clamping fixture that restrain the outer arm  120  in a position that resembles its assembled state as part of rocker arm assembly  100 . Grinding of these surfaces requires that the pads  130 ,  132  remain parallel to one another and that the outer arm  120  not be distorted. Clamping at the clamping lobes  150  prevents distortion that may occur to the outer arm  120  under other clamping arrangements. For example, clamping at the clamping lobes  150 , which are preferably integral to the outer arm  120 , assist in eliminating any mechanical stress that may occur by clamping that squeezes outer side arms  124 ,  126  toward one another. In another example, the location of clamping lobe  150  immediately underneath slider pads  130 ,  132 , results in substantially zero to minimal torque on the outer arm  120  caused by contact forces with the grinding machine. In certain applications, it may be necessary to apply pressure to other portions in outer arm  120  in order to minimize distortion. 
     4.5 DVVL Assembly Operation 
       FIG. 19  illustrates an exploded view of the switching rocker arm  100  of  FIGS. 27 and 15 . With reference to  FIGS. 19 and 28 , when assembled, roller bearing  128  is part of a needle roller-type assembly  129 , which may have needles  180  mounted between the roller bearing  128  and roller axle  182 . Roller axle  182  is mounted to the inner arm  122  via roller axle apertures  183 ,  184 . Roller assembly  129  serves to transfer the rotational motion of the low-lift cam  108  to the inner rocker arm  122 , and in turn transfer motion to the valve  112  in the unlatched state. Pivot axle  118  is mounted to inner arm  122  through collar  123  and to outer arm  120  through pivot axle apertures  160 ,  162  at the first end  101  of rocker arm  100 . Lost motion rotation of the outer arm  120  relative to the inner arm  122  in the unlatched state occurs about pivot axle  118 . Lost motion movement in this context means movement of the outer arm  120  relative to the inner arm  122  in the unlatched state. This motion does not transmit the rotating motion of the first and second high-lift lobe  104 ,  106  of the cam  102  to the valve  112  in the unlatched state. 
     Other configurations other than the roller assembly  129  and pads  130 ,  132  also permit the transfer of motion from cam  102  to rocker arm  100 . For example, a smooth non-rotating surface (not shown) such as pads  130 ,  132  may be placed on inner arm  122  to engage low-lift lobe  108 , and roller assemblies may be mounted to rocker arm  100  to transfer motion from high-lift lobes  104 ,  106  to outer arm  120  of rocker arm  100 . 
     Now, with reference to  FIGS. 4, 19, and 12 , as noted above, the exemplary switching rocker arm  100  uses a three-lobed cam  102 . 
     To make the design compact, with dynamic loading as close as possible to non-switching rocker arm designs, slider pads  130 ,  132  are used as the surfaces that contact the cam lobes  104 ,  106  during operation in high-lift mode. Slider pads produce more friction during operation than other designs such as roller bearings, and the friction between the first slider pad surface  130  and the first high-lift lobe surface  104 , plus the friction between the second slider pad  132  and the second high-lift lobe  106 , creates engine efficiency losses. 
     When the rocker arm assembly  100  is in high-lift mode, the full load of the valve opening event is applied slider pads  130 ,  132 . When the rocker arm assembly  100  is in low-lift mode, the load of the valve opening event applied to slider pads  130 ,  132  is less, but present. Packaging constraints for the exemplary switching rocker arm  100 , require that the width of each slider pad  130 ,  132  as described by slider pad edge length  710 ,  711  that come in contact with the cam lobes  104 ,  106  are narrower than most existing slider interface designs. This results in higher part loading and stresses than most existing slider pad interface designs. The friction results in excessive wear to cam lobes  104 ,  106 , and slider pads  130 ,  132 , and when combined with higher loading, may result in premature part failure. In the exemplary switching rocker arm assembly, a coating such as a diamond like carbon coating is used on the slider pads  130 ,  132  on the outer arm  120 . 
     A diamond-like carbon coating (DLC) coating enables operation of the exemplary switching rocker arm  100  by reducing friction, and at the same providing necessary wear and loading characteristics for the slider pad surfaces  130 ,  132 . As can be easily seen, benefits of DLC coating can be applied to any part surfaces in this assembly or other assemblies, for example the pivot axle surfaces  160 ,  162 , on the outer arm  120  described in  FIG. 19 . 
     Although similar coating materials and processes exist, none are sufficient to meet the following DVVL rocker arm assembly requirements: 1) be of sufficient hardness, 2) have suitable loadbearing capacity, 3) be chemically stable in the operating environment, 4) be applied in a process where temperatures do not exceed the annealing temperature for the outer arm  120 , 5) meet engine lifetime requirements, and 6) offer reduced friction as compared to a steel on steel interface. The DLC coating process described earlier meets the requirements set forth above, and is applied to slider pad surfaces  130 ,  132 , which are ground to a final finish using a grinding wheel material and speed that is developed for DLC coating applications. The slider pad surfaces  130 ,  132  are also polished to a specific surface roughness, applied using one of several techniques, for example vapor honing or fine particle sand blasting. 
     4.5.1 Hydraulic Fluid System 
     The hydraulic latch for rocker arm assembly  100  must be built to fit into a compact space, meet switching response time requirements, and minimize oil pumping losses. Oil is conducted along fluid pathways at a controlled pressure, and applied to controlled volumes in a way that provides the necessary force and speed to activate latch pin switching. The hydraulic conduits require specific clearances, and sizes so that the system has the correct hydraulic stiffness and resulting switching response time. The design of the hydraulic system must be coordinated with other elements that comprise the switching mechanism, for example the biasing spring  230 . 
     In the switching rocker arm  100 , oil is transmitted through a series of fluid-connected chambers and passages to the latch pin assembly  201 , or any other hydraulically activated latch pin mechanism. As described above, the hydraulic transmission system begins at oil flow port  506  in the DFHLA  110 , where oil or another hydraulic fluid at a controlled pressure is introduced. Pressure can be modulated with a switching device, for example, a solenoid valve. After leaving the ball plunger end  601 , oil or other pressurized fluid is directed from this single location, through the first oil gallery  144  and the second oil gallery  146  of the inner arm discussed above, which have bores sized to minimize pressure drop as oil flows from the ball socket  502 , shown in  FIG. 10 , to the latch pin assembly  201  in  FIG. 19 . 
     The latch pin assembly  201  for latching inner arm  122  to outer arm  120 , which in the illustrated embodiment is found near second end  103  of rocker arm  100 , is shown in  FIG. 19  as including a latch pin  200  that is extended in high-lift mode, securing inner arm  122  to outer arm  120 . In low-lift mode, latch  200  is retracted into inner arm  122 , allowing lost motion movement of outer arm  120 . Oil pressure is used to control latch pin  200  movement. 
     As illustrated in  FIG. 32 , one embodiment of a latch pin assembly shows that the oil galleries  144 ,  146  (shown in  FIG. 19 ) are in fluid communication with the chamber  250  through oil opening  280 . 
     The oil is provided to oil opening  280  and the latch pin assembly  201  at a range of pressures, depending on the required mode of operation. 
     As can be seen in  FIG. 33 , upon introduction of pressurized oil into chamber  250 , latch  200  retracts into bore  240 , allowing outer arm  120  to undergo lost motion rotation with respect to inner arm  122 . Oil can be transmitted between the first generally cylindrical surface  205  and surface  241 , from first chamber  250  to second chamber  425  shown in  FIG. 32 . 
     Some of the oil exits back to the engine through hole  209 , drilled into the inner arm  122 . The remaining oil is pushed back through the hydraulic pathways as the biasing spring  230  expands when it returns to the latched high-lift state. It can be seen that a similar flow path can be employed for latch mechanisms that are biased for normally unlatched operation. 
     The latch pin assembly design manages latch pin response time through a combination of clearances, tolerances, hole sizes, chamber sizes, spring designs, and similar metrics that control the flow of oil. For example, the latch pin design may include features such as a dual diameter pin designed with an active hydraulic area to operate within tolerance in a given pressure range, an oil sealing land designed to limit oil pumping losses, or a chamfer oil in-feed. 
     Now, with reference to  FIGS. 32-34 , latch  200  contains design features that provide multiple functions in a limited space:
         1. Latch  200  employs the first generally cylindrical surface  205  and the second generally cylindrical surface  206 . First generally cylindrical surface  205  has a diameter larger than that of the second generally cylindrical surface  206 . When pin  200  and sleeve  210  are assembled together in bore  240 , a chamber  250  is formed without employing any additional parts. As noted, this volume is in fluid communication with oil opening  280 . Additionally, the area of pressurizing surface  422 , combined with the transmitted oil pressure, can be controlled to provide the necessary force to move the pin  200 , compress the biasing spring  230 , and switch to low-lift mode (unlatched).   2. The space between the first generally cylindrical surface  205  and the adjacent bore wall  241  is intended to minimize the amount of oil that flows from chamber  250  into second chamber  425 . The clearance between the first generally cylindrical surface  205  and surface  241  must be closely controlled to allow freedom of movement of pin  200  without oil leakage and associated oil pumping losses as oil is transmitted between first generally cylindrical surface  205  and surface  241 , from chamber  250  to second chamber  425 .   3. Package constraints require that the distance along the axis of movement of the pin  200  be minimized. In some operating conditions, the available oil sealing land  424  may not be sufficient to control the flow of oil that is transmitted between first generally cylindrical surface  205  and surface  241 , from chamber  250  to the second chamber  425 . An annular sealing surface is described. As latch  200  retracts, it encounters bore wall  208  with its rear surface  203 . In one preferred embodiment, rear surface  203  of latch  200  has a flat annular or sealing surface  207  that lies generally perpendicular to first and second generally cylindrical bore wall  241 ,  242 , and parallel to bore wall  208 . The flat annular surface  207  forms a seal against bore wall  208 , which reduces oil leakage from chamber  250  through the seal formed by first generally cylindrical surface  205  of latch  200  and first generally cylindrical bore wall  241 . The area of sealing surface  207  is sized to minimize separation resistance caused by a thin film of oil between the sealing surface  207  and the bore wall  208  shown in  FIG. 32 , while maintaining a seal that prevents pressurized oil from flowing between the sealing surface  207  and the bore wall  208 , and out hole  209 .   4. In one latch pin  200  embodiment, an oil in-feed surface  426 , for example a chamfer, provides an initial pressurizing surface area to allow faster initiation of switching, and overcome separation resistance caused by a thin film of oil between the pressurization surface  422  and the sleeve end  427 . The size and angle of the chamfer allows ease of switching initiation, without unplanned initiation due to oil pressure variations encountered during normal operation. In a second latch pin  200  embodiment, a series of castellations  432 , arranged radially as shown in  FIG. 34 , provide an initial pressurizing surface area, sized to allow faster initiation of switching, and overcome separation resistance caused by a thin film of oil between the pressurization surface  422  and the sleeve end  427 .       

     An oil in-feed surface  426  can also reduce the pressure and oil pumping losses required for switching by lowering the requirement for the breakaway force between pressurization surface  422  and the sleeve end  427 . These relationships can be shown as incremental improvements to switching response and pumping losses. 
     As oil flows throughout the previously-described switching rocker arm assembly  100  hydraulic system, the relationship between oil pressure and oil fluid pathway area and length largely defines the reaction time of the hydraulic system, which also directly affects switching response time. For example, if high pressure oil at high velocity enters a large volume, its velocity will suddenly slow, decreasing its hydraulic reaction time, or stiffness. A range of these relationships that are specific to the operation of switching rocker arm assembly  100  can be calculated. One relationship, for example, can be described as follows: oil at a pressure of 2 bar is supplied to chamber  250 , where the oil pressure, divided by the pressurizing surface area, transmits a force that overcomes biasing spring  230  force, and initiates switching within 10 milliseconds from latched to unlatched operation. 
     A range of characteristic relationships that result in acceptable hydraulic stiffness and response time, with minimized oil pumping losses can be calculated from system design variables that can be defined as follows:
         Oil gallery  144 ,  146  inside diameter and length from the ball socket  502  to hole  280 .   Bore hole  280  diameter and length   Area of pressurizing surface  422     The volume of chamber  250  in all states of operation   The volume of second chamber  425  in all states of operation   Cross-sectional area created by the space between first generally cylindrical surface  205  and surface  241 .   The length of oil sealing land  424     The area of the flat annular surface  207     The diameter of hole  209     Oil pressure supplied by the DFHLA  110     Stiffness of biasing spring  230     The cross sectional area and length of flow channels  504 ,  508 ,  509     The area and number of oil in-feed surfaces  426 .   The number and cross sectional area of castellations  432         

     Latch response times for the previously described hydraulic arrangement in switching rocker arm  100  can be described for a range of conditions, for example: 
     Oil temperatures: 10° C. to 120° C. 
     Oil type: 5w-20 weight 
     This conditions result in a range of oil viscosities that affect the latch response time. 
     4.5.2 Latch Pin Mechanism 
     The latch pin mechanism  201  of rocker arm assembly  100  provides a means of mechanically switching from high-lift to low-lift and vice versa. A latch pin mechanism can be configured to be normally in an unlatched or latched state. Several preferred embodiments can be described. 
     In one embodiment, the latch pin assembly  201  for latching inner arm  122  to outer arm  120 , which is found near second end  103  of rocker arm  100 , is shown in  FIG. 19  as comprising latch pin  200 , sleeve  210 , orientation pin  220 , and latch spring  230 . The latch pin assembly  201  is configured to be mounted inside inner arm  122  within bore  240 . As explained below, in the assembled rocker arm  100 , latch  200  is extended in high-lift mode, securing inner arm  122  to outer arm  120 . In low-lift mode, latch  200  is retracted into inner arm  122 , allowing lost motion movement of outer arm  120 . Switched oil pressure, as described previously, is provided through the first and second oil gallery  144 ,  146  to control whether latch  200  is latched or unlatched. Plugs  170  are inserted into gallery holes  172  to form a pressure tight seal closing first and second oil gallery  144 ,  146  and allowing them to pass oil to latching mechanism  201 . 
       FIG. 32  illustrates a cross-sectional view of the latching mechanism  201  in its latched state along the line  32 ,  33 - 32 ,  33  in  FIG. 28 . A latch  200  is disposed within bore  240 . Latch  200  has a spring bore  202  in which biasing spring  230  is inserted. The latch  200  has a rear surface  203  and a front surface  204 . Latch  200  also employs the first generally cylindrical surface  205  and a second generally cylindrical surface  206 . First generally cylindrical surface  205  has a diameter larger than that of the second generally cylindrical surface  206 . Spring bore  202  is generally concentric with surfaces  205 ,  206 . 
     Sleeve  210  has a generally cylindrical outer surface  211  that interfaces a first generally cylindrical bore wall  241 , and a generally cylindrical inner surface  215 . Bore  240  has a first generally cylindrical bore wall  241 , and a second generally cylindrical bore wall  242  having a larger diameter than first generally cylindrical bore wall  241 . The generally cylindrical outer surface  211  of sleeve  210  and first generally cylindrical surface  205  of latch  200  engage first generally cylindrical bore wall  241  to form tight pressure seals. Further, the generally cylindrical inner surface  215  of sleeve  210  also forms a tight pressure seal with second generally cylindrical surface  206  of latch  200 . During operation, these seals allow oil pressure to build in chamber  250 , which encircles second generally cylindrical surface  206  of latch  200 . 
     The default position of latch  200 , shown in  FIG. 32 , is the latched position. Spring  230  biases latch  200  outwardly from bore  240  into the latched position. Oil pressure applied to chamber  250  retracts latch  200  and moves it into the unlatched position. Other configurations are also possible, such as where spring  230  biases latch  200  in the unlatched position, and application of oil pressure between bore wall  208  and rear surface  203  causes latch  200  to extend outwardly from the bore  240  to latch outer arm  120 . 
     In the latched state, latch  200  engages a latch surface  214  of outer arm  120  with arm engaging surface  213 . As shown in  FIG. 32 , outer arm  120  is impeded from moving downward and will transfer motion to inner arm  122  through latch  200 . An orientation feature  212  takes the form of a channel into which orientation pin  221  extends from outside inner arm  122  through first pin opening  217  and then through second pin opening  218  in sleeve  210 . The orientation pin  221  is generally solid and smooth. A retainer  222  secures pin  221  in place. The orientation pin  221  prevents excessive rotation of latch  200  within bore  240 . 
     As previously described, and seen in  FIG. 33 , upon introduction of pressurized oil into chamber  250 , latch  200  retracts into bore  240 , allowing outer arm  120  to undergo lost motion rotation with respect to inner arm  122 . The outer arm  120  is then no longer impeded by latch  200  from moving downward and exhibiting lost motion movement. Pressurized oil is introduced into chamber  250  through oil opening  280 , which is in fluid communication with oil galleries  144 ,  146 . 
       FIGS. 35A-35F  illustrate several retention devices for orientation pin  221 . In  FIG. 35A , pin  221  is cylindrical with a uniform thickness. A push-on ring  910 , as shown in  FIG. 35C  is located in recess  224  located in sleeve  210 . Pin  221  is inserted into ring  910 , causing teeth  912  to deform and secure pin  221  to ring  910 . Pin  221  is then secured in place due to the ring  910  being enclosed within recess  224  by inner arm  122 . In another embodiment, shown in  FIG. 35B , pin  221  has a slot  902  in which teeth  912  of ring  910  press, securing ring  910  to pin  221 . In another embodiment shown in  FIG. 35D , pin  221  has a slot  904  in which an E-styled clip  914  of the kind shown in  FIG. 35E , or a bowed E-styled clip  914  as shown in  FIG. 35F  may be inserted to secure pin  221  in place with respect to inner arm  122 . In yet other embodiments, wire rings may be used in lieu of stamped rings. During assembly, the E-styled clip  914  is placed in recess  224 , at which point the sleeve  210  is inserted into inner arm  122 , then, the orientation pin  221  is inserted through the clip  910 . 
     An exemplary latch  200  is shown in  FIG. 36 . The latch  200  is generally divided into a head portion  290  and a body portion  292 . The front surface  204  is a protruding convex curved surface. This surface shape extends toward outer arm  120  and results in an increased chance of proper engagement of arm engaging surface  213  of latch  200  with outer arm  120 . Arm engaging surface  213  comprises a generally flat surface. Arm engaging surface  213  extends from a first boundary  285  with second generally cylindrical surface  206  to a second boundary  286  and from a boundary  287  with the front surface to a boundary  233  with surface  232 . The portion of arm engaging surface  213  that extends furthest from surface  232  in the direction of the longitudinal axis A of latch  200  is located substantially equidistant between first boundary  285  and second boundary  286 . Conversely, the portion of arm engaging surface  213  that extends the least from surface  232  in the axial direction A is located substantially at first and second boundaries  285 ,  286 . Front surface  204  need not be a convex curved surface but instead can be a v-shaped surface, or some other shape. The arrangement permits greater rotation of the latch  200  within bore  240  while improving the likelihood of proper engagement of arm engaging surface  213  of latch  200  with outer arm  120 . 
     An alternative latching mechanism  201  is shown in  FIG. 37 . An orientation plug  1000 , in the form of a hollow cup-shaped plug, is press-fit into sleeve hole  1009  and orients latch  200  by extending into orientation feature  212 , preventing latch  200  from rotating excessively with respect to sleeve  210 . As discussed further below, an aligning slot  1004  assists in orienting the latch  200  within sleeve  210  and ultimately within inner arm  122  by providing a feature by which latch  200  may be rotated within the sleeve  210 . The alignment slot  1004  may serve as a feature with which to rotate the latch  200 , and also to measure its relative orientation. 
     With reference to  FIGS. 38-40 , an exemplary method of assembling a switching rocker arm  100  is as follows: the orientation plug  1000  is press-fit into sleeve hole  1009  and latch  200  is inserted into generally cylindrical inner surface  215  of sleeve  210 . 
     The latch pin  200  is then rotated clockwise until orientation feature  212  reaches plug  1000 , at which point interference between the orientation feature  212  and plug  1000  prevents further rotation. An angle measurement A 1 , as shown in  FIG. 38 , is then taken corresponding to the angle between arm engaging surface  213  and sleeve references  1010 ,  1012 , which are aligned to be perpendicular to sleeve hole  1009 . Aligning slot  1004  may also serve as a reference line for latch  200 , and key slots  1014  may also serve as references located on sleeve  210 . The latch pin  200  is then rotated counterclockwise until orientation feature  212  reaches plug  1000 , preventing further rotation. As seen in  FIG. 39 , a second angle measurement A 2  is taken corresponding to the angle between arm engaging surface  213  and sleeve references  1010 ,  1012 . Rotating counterclockwise and then clockwise is also permissible in order to obtain A 1  and A 2 . As shown in  FIG. 40 , upon insertion into the inner arm  122 , the sleeve  210  and pin subassembly  1200  is rotated by an angle A as measured between inner arm references  1020  and sleeve references  1010 ,  1012 , resulting in the arm engaging surface  213  being oriented horizontally with respect to inner arm  122 , as indicated by inner arm references  1020 . The amount of rotation A should be chosen to maximize the likelihood the latch  200  will engage outer arm  120 . One such example is to rotate subassembly  1200  an angle half of the difference of A 2  and A 1  as measured from inner arm references  1020 . Other amounts of adjustment A are possible within the scope of the present disclosure. 
     A profile of an alternative embodiment of pin  1000  is shown in  FIG. 41 . Here, the pin  1000  is hollow, partially enclosing an inner volume  1050 . The pin has a substantially cylindrical first wall  1030  and a substantially cylindrical second wall  1040 . The substantially cylindrical first wall  1030  has a diameter D 1  larger than diameter D 2  of second wall  1040 . In one embodiment shown in  FIG. 41 , a flange  1025  is used to limit movement of pin  1000  downwardly through pin opening  218  in sleeve  210 . In a second embodiment shown in  FIG. 42 , a press-fit limits movement of pin  1000  downwardly through pin opening  218  in sleeve  210 . 
     4.6 DVVL Assembly Lash Management 
     A method of managing three or more lash values, or design clearances, in the DVVL switching rocker arm assembly  100  shown in  FIG. 4 , is described. Methods may include a range of manufacturing tolerances, wear allowances, and design profiles for cam lobe/rocker arm contact surfaces. 
     DVVL Assembly Lash Description 
     An exemplary rocker arm assembly  100  shown in  FIG. 4  has one or more lash values that must be maintained in one or more locations in the assembly. The three-lobed cam  102 , illustrated in  FIG. 4 , is comprised of three cam lobes, a first high lift lobe  104 , a second high lift lobe  106 , and a low lift lobe  108 . Cam lobes  104 ,  106 , and  108 , are comprised of profiles that respectively include a base circle  605 ,  607 ,  609 , described as generally circular and concentric with the cam shaft. 
     The switching rocker arm assembly  100  shown in  FIG. 4  was designed to have small clearances (lash) in two locations. The first location, illustrated in  FIG. 43 , is latch lash  602 , the distance between latch pad surface  214  and the arm engaging surface  213 . Latch lash  602  ensures that the latch  200  is not loaded and can move freely when switching between high-lift and low-lift modes. As shown in  FIGS. 4, 27, 43, and 49 , a second example of lash, the distance between the first slider pad  130  and the first high lift cam lobe base circle  605 , is illustrated as camshaft lash  610 . Camshaft lash  610  eliminates contact, and by extension, friction losses, between slider pads  130 ,  132 , and their respective high lift cam lobe base circles  605 ,  607  when the roller bearing  128 , shown in  FIG. 49 , is contacting the low-lift cam base circle  609  during low-lift operation. 
     During low-lift mode, camshaft lash  610  also prevents the torsion spring  134 ,  136  force from being transferred to the DFHLA  110  during base circle  609  operation. This allows the DFHLA  110  to operate like a standard rocker arm assembly with normal hydraulic lash compensation where the lash compensation portion of the DFHLA is supplied directly from an engine oil pressure gallery. As shown in  FIG. 47 , this action is facilitated by the rotational stop  621 ,  623  within the switching rocker arm assembly  100  that prevents the outer arm  120  from rotating sufficiently far due to the torsion spring  134 ,  136  force to contact the high lift lobes  104 ,  106 . 
     As illustrated in  FIGS. 43 and 48 , total mechanical lash is the sum of camshaft lash  610  and latch lash  602 . The sum affects valve motion. The high lift camshaft profiles include opening and closing ramps  661  to compensate for total mechanical lash  612 . Minimal variation in total mechanical lash  612  is important to maintain performance targets throughout the life of the engine. To keep lash within the specified range, the total mechanical lash  612  tolerance is closely controlled in production. Because component wear correlates to a change in total mechanical lash, low levels of component wear are allowed throughout the life of the mechanism. Extensive durability shows that allocated wear allowance and total mechanical lash remain within the specified limits through end of life testing. 
     Referring to the graph shown in  FIG. 48 , lash in millimeters is on the vertical axis, and camshaft angle in degrees is arranged on the horizontal axis. The linear portion  661  of the valve lift profile  660  shows a constant change of distance in millimeters for a given change in camshaft angle, and represents a region where closing velocity between contact surfaces is constant. For example, during the linear portion  661  of the valve lift profile curve  660 , when the rocker arm assembly  100  ( FIG. 4 ) switches from low-lift mode to high-lift mode, the closing distance between the first slider pad  130 , and the first high-lift lobe  104  ( FIG. 43 ), represents a constant velocity. Utilizing the constant velocity region reduces impact loading due to acceleration. 
     As noted in  FIG. 48 , no valve lift occurs during the constant velocity no lift portion  661  of the valve lift profile curve  660 . If total lash is reduced or closely controlled through improved system design, manufacturing, or assembly processes, the amount of time required for the linear velocity portion of the valve lift profile is reduced, providing engine management benefits, for example allowing earlier valve lift opening or consistent valve operation engine to engine. 
     Now, as to  FIGS. 43, 47, and 48 , design and assembly variations for individual parts and sub-assemblies can produce a matrix of lash values that meet switch timing specifications and reduce the required constant velocity switching region described previously. For example, one latch pin  200  self-aligning embodiment may include a feature that requires a minimum latch lash  602  of 10 microns to function. An improved modified latch  200 , configured without a self-aligning feature may be designed that requires a latch lash  602  of 5 microns. This design change decreases the total lash by 5 microns, and decreases the required no lift  661  portion of the valve lift profile  660 . 
     Latch lash  602 , and camshaft lash  610  shown in  FIG. 43 , can be described in a similar manner for any design variation of switching rocker arm assembly  100  of  FIG. 4  that uses other methods of contact with the three-lobed cam  102 . In one embodiment, a sliding pad similar to  130  is used instead of roller bearing  128  ( FIGS. 15 and 27 ). In a second embodiment, rollers similar to  128  are used in place of slider pad  130  and slider pad  132 . There are also other embodiments that have combinations of rollers and sliders. 
     Lash Management, Testing 
     As described in following sections, the design and manufacturing methods used to manage lash were tested and verified for a range of expected operating conditions to simulate both normal operation and conditions representing higher stress conditions. 
     Durability of the DVVL switching rocker arm is assessed by demonstrating continued performance (i.e., valves opening and closing properly) combined with wear measurements. Wear is assessed by quantifying loss of material on the DVVL switching rocker arm, specifically the DLC coating, along with the relative amounts of mechanical lash in the system. As noted above, latch lash  602  ( FIG. 43 ) is necessary to allow movement of the latch pin between the inner and outer arm to enable both high and low lift operation when commanded by the engine electronic control unit (ECU). An increase in lash for any reason on the DVVL switching rocker arm reduces the available no-lift ramp  661  ( FIG. 48 ), resulting in high accelerations of the valve-train. The specification for wear with regards to mechanical lash is set to allow limit build parts to maintain desirable dynamic performance at end of life. 
     For example, as shown in  FIG. 43 , wear between contacting surfaces in the rocker arm assembly will change latch lash  602 , cam shaft lash  610 , and the resulting total lash. Wear that affects these respective values can be described as follows: 1) wear at the interface between the roller bearing  128  ( FIG. 15 ) and the cam lobe  108  ( FIG. 4 ) reduces total lash, 2) wear at the sliding interface between slider pads  130 ,  132  ( FIG. 15 ) and cam lobes  104 ,  106  ( FIG. 4 ) increases total lash, and 3) wear between the latch  200  and the latch pad surface  214  increases total lash. Since bearing interface wear decreases total lash and latch and slider interface wear increase total lash, overall wear may result in minimal net total lash change over the life of the rocker arm assembly. 
     4.7 DVVL Assembly Dynamics 
     The weight distribution, stiffness, and inertia for traditional rocker arms have been optimized for a specified range of operating speeds and reaction forces that are related to dynamic stability, valve tip loading and valve spring compression during operation. An exemplary switching rocker arm  100 , illustrated in  FIG. 4  has the same design requirements as the traditional rocker arm, with additional constraints imposed by the added mass and the switching functions of the assembly. Other factors must be considered as well, including shock loading due to mode-switching errors and subassembly functional requirements. Designs that reduce mass and inertia, but do not effectively address the distribution of material needed to maintain structural stiffness and resist stress in key areas can result in parts that deflect out of specification or become overstressed, both of which are conditions that may lead to poor switching performance and premature part failure. The DVVL rocker arm assembly  100 , shown in  FIG. 4 , must be dynamically stable to 3500 rpm in low lift mode and 7300 rpm in high lift mode to meet performance requirements. 
     As to  FIGS. 4, 15, 19, and 27 , DVVL rocker arm assembly  100  stiffness is evaluated in both low lift and high lift modes. In low lift mode, the inner arm  122  transmits force to open the valve  112 . The engine packaging volume allowance and the functional parameters of the inner arm  122  do not require a highly optimized structure, as the inner arm stiffness is greater than that of a fixed rocker arm for the same application. In high lift mode, the outer arm  120  works in conjunction with the inner arm  122  to transmit force to open the valve  112 . Finite Element Analysis (FEA) techniques show that the outer arm  120  is the most compliant member, as illustrated in  FIG. 50  in an exemplary plot showing a maximum area of vertical deflection  670 . Mass distribution and stiffness optimization for this part is focused on increasing the vertical section height of the outer arm  120  between the slider pads  130 ,  132  and the latch  200 . Design limits on the upper profile of the outer arm  120  are based on clearance between the outer arm  120  and the swept profile of the high lift lobes  104 ,  106 . Design limits on the lower profile of the outer arm  120  are based on clearance to the valve spring retainer  116  in low lift mode. Optimizing material distribution within the described design constraints decreases the vertical deflection and increased stiffness, in one example, more than 33 percent over initial designs. 
     As shown in  FIGS. 15 and 52 , the DVVL rocker arm assembly  100  is designed to minimize inertia as it pivots about the ball plunger contact point  611  of the DFHLA  110  by biasing mass of the assembly as much as possible towards side  101 . This results in a general arrangement with two components of significant mass, the pivot axle  118  and the torsion springs  134   136 , located near the DFHLA  110  at side  101 . With pivot axle  118  in this location, the latch  200  is located at end  103  of the DVVL rocker arm assembly  100 . 
       FIG. 55  is a plot that compares the DVVL rocker arm assembly  100  stiffness in high-lift mode with other standard rocker arms. The DVVL rocker arm assembly  100  has lower stiffness than the fixed rocker arm for this application; however, its stiffness is in the existing range rocker arms used in similar valve train configurations now in production. The inertia of the DVVL rocker arm assembly  100  is approximately double the inertia of a fixed rocker arm, however, its inertia is only slightly above the mean for rocker arms used in similar valve train configurations now in production. The overall effective mass of the intake valve train, consisting of multiple DVVL rocker arm assemblies  100  is 28% greater than a fixed intake valve train. These stiffness, mass, and inertia values require optimization of each component and subassembly to ensure minimum inertia and maximum stiffness while meeting operational design criteria. 
     4.7.1 DVVL Assembly Dynamics Detailed Description 
     The major components that comprise total inertia for the rocker arm assembly  100  are illustrated in  FIG. 53 . These are the inner arm assembly  622 , the outer arm  120 , and the torsion springs  134 ,  136 . As noted, functional requirements of the inner arm assembly  622 , for example, its hydraulic fluid transfer pathways and its latch pin mechanism housing, require a stiffer structure than a fixed rocker arm for the same application. In the following description, the inner arm assembly  622  is considered a single part. 
     Referring to  FIGS. 51-53 ,  FIG. 51  shows a top view of the rocker arm assembly  100  in  FIG. 4 .  FIG. 52  is a section view along the line  52 - 52  in  FIG. 51  that illustrates loading contact points for the rocker arm assembly  100 . The rotating three lobed cam  102  imparts a cam load  616  to the roller bearing  128  or, depending on mode of operation, to the slider pads  130 ,  132 . The ball plunger end  601  and the valve tip  613  provide opposing forces. 
     In low-lift mode, the inner arm assembly  622  transmits the cam load  616  to the valve tip  613 , compresses spring  114  (of  FIG. 4 ), and opens the valve  112 . In high-lift mode, the outer arm  120 , and the inner arm assembly  622  are latched together. In this case, the outer arm  120  transmits the cam load  616  to the valve tip  613 , compresses the spring  114 , and opens the valve  112 . 
     Now, as to  FIGS. 4 and 52 , the total inertia for the rocker arm assembly  100  is determined by the sum of the inertia of its major components, calculated as they rotate about the ball plunger contact point  611 . In the exemplary rocker arm assembly  100 , the major components may be defined as the torsion springs  134 ,  136 , the inner arm assembly  622 , and the outer arm  120 . When the total inertia increases, the dynamic loading on the valve tip  613  increases, and system dynamic stability decreases. To minimize valve tip loading and maximize dynamic stability, mass of the overall rocker arm assembly  100  is biased towards the ball plunger contact point  611 . The amount of mass that can be biased is limited by the required stiffness of the rocker arm assembly  100  needed for a given cam load  616 , valve tip load  614 , and ball plunger load  615 . 
     Now, as to  FIGS. 4 and 52 , the stiffness of the rocker arm assembly  100  is determined by the combined stiffness of the inner arm assembly  622 , and the outer arm  120 , when they are in a high-lift or low-lift state. Stiffness values for any given location on the rocker arm assembly  100  can be calculated and visualized using Finite Element Analysis (FEA) or other analytical methods, and characterized in a plot of stiffness versus location along the measuring axis  618 . In a similar manner, stiffness for the outer arm  120  and inner arm assembly  622  can be individually calculated and visualized using Finite Element Analysis (FEA) or other analytical methods. An exemplary illustration  106  shows the results of these analyses as a series characteristic plots of stiffness versus location along the measuring axis  618 . As an additional illustration noted earlier,  FIG. 50  illustrates a plot of maximum deflection for the outer arm  120 . 
     Now, referencing  FIGS. 52 and 56 , stress and deflection for any given location on the rocker arm assembly  100  can be calculated using Finite Element Analysis (FEA) or other analytical methods, and characterized as plots of stress and deflection versus location along the measuring axis  618  for given cam load  616 , valve tip load  614 , and ball plunger load  615 . In a similar manner, stress and deflection for the outer arm  120  and inner arm assembly  622  can be individually calculated and visualized using Finite Element Analysis (FEA) or other analytical methods. An exemplary illustration in  FIG. 56  shows the results of these analyses as a series of characteristic plots of stress and deflection versus location along the measuring axis  618  for given cam load  616 , valve tip load  614 , and ball plunger load  615 . 
     4.7.2 DVVL Assembly Dynamics Analysis 
     For stress and deflection analysis, a load case is described in terms of load location and magnitude as illustrated in  FIG. 52 . For example, in a latched rocker arm assembly  100  in high-lift mode, the cam load  616  is applied to slider pads  130 ,  132 . The cam load  616  is opposed by the valve tip load  614  and the ball plunger load  615 . The first distance  632  is the distance measured along the measuring axis  618  between the valve tip load  614  and the ball plunger load  615 . The second distance  634  is the distance measured along the measuring axis  618  between the valve tip load  614  and the cam load  616 . The load ratio is the second distance  634  divided by the first distance  632 . For dynamic analysis, multiple values and operating conditions are considered for analysis and possible optimization. These may include the three lobe camshaft interface parameters, torsion spring parameters, total mechanical lash, inertia, valve spring parameters, and DFHLA parameters. 
     Design parameters for evaluation can be described: 
     
       
         
           
               
               
               
             
               
                   
               
               
                 Variable/ 
                   
                 Value/Range for a Design 
               
               
                 Parameter 
                 Description 
                 Iteration 
               
               
                   
               
             
            
               
                 Engine 
                 The maximum rotational speed of the rocker arm 
                 7300 rpm in high-lift mode 
               
               
                 speed 
                 assembly 100 about the ball plunger contact point 
                 3500 rpm in low-lift mode 
               
               
                   
                 611 is derived from the engine speed. 
                   
               
               
                 Lash 
                 Lash enables switching from between high-lift and 
                 Cam lash 
               
               
                   
                 low-lift modes, and varies based on the selected 
                 Latch lash 
               
               
                   
                 design. In the example configuration shown in 
                 Total lash 
               
               
                   
                 FIG. 52, a deflection of the outer arm 120 slider 
                   
               
               
                   
                 pad results in a decrease of the total lash available 
                   
               
               
                   
                 for switching. 
                   
               
               
                 Maximum 
                 This value is based on the selected design 
                 Total lash +/− tolerance 
               
               
                 allowable 
                 configuration. 
                   
               
               
                 deflection 
                   
                   
               
               
                 Maximum 
                 Establish allowable loading for the specified 
                 Kinematic contact stresses: 
               
               
                 allowable 
                 materials of construction. 
                 Valve tip = 
               
               
                 stress 
                   
                 Ball plunger end = 
               
               
                   
                   
                 Roller = 1200-1400 MPa 
               
               
                   
                   
                 Slider pads = 800-1000 MPa 
               
               
                 Dynamic 
                   
                 Valve closing velocity 
               
               
                 stability 
                   
                   
               
               
                 Cam shape 
                 The cam load 616 in FIG. 52 is established by the 
                 This variable is considered fixed 
               
               
                   
                 rotating cam lobe as it acts to open the valve. The 
                 for iterative design analysis. 
               
               
                   
                 shape of the cam lobe affects dynamic loading. 
                   
               
               
                 Valve spring 
                 The spring 114 compression stiffness is fixed for a 
                   
               
               
                 stiffness 
                 given engine design. 
                   
               
               
                 Ball plunger 
                 As described in FIG. 52, the second distance 632 
                 Range = 20-50 mm 
               
               
                 to valve tip 
                 value is set by the engine design. 
                   
               
               
                 distance 
                   
                   
               
               
                 Load ratio 
                 The load ratio as shown in FIG. 52 is the second 
                 Range = 0.2-0.8 
               
               
                   
                 distance 634 divided by the first distance 632. This 
                   
               
               
                   
                 value is imposed by the design configuration and 
                   
               
               
                   
                 load case selected. 
                   
               
               
                 Inertia 
                 This is a calculated value. 
                 Range = 20-60 Kg*mm2 
               
               
                   
               
            
           
         
       
     
     Now, as referenced by  FIGS. 4, 51, 52, 53, and 54 , based on given set of design parameters, a general design methodology is described.
         1. In step one  350 , arrange components  622 ,  120 ,  134 , and  136  along the measuring axis to bias mass towards the ball plunger contact point  611 . For example, the torsion springs  134 ,  136  may be positioned 2 mm to the left of the ball plunger contact point, and the pivot axle  118  in the inner arm assembly  622  may be positioned 5 mm to the right. The outer arm  120  is positioned to align with the pivot axle  118  as shown in  FIG. 53 .   2. In step  351 , for a given component arrangement, calculate the total inertia for the rocker arm assembly  100 .   3. In step  352 , evaluate the functionality of the component arrangement. For example, confirm that the torsion springs  134 ,  136  can provide the required stiffness in their specified location to keep the slider pads  130 ,  132  in contact with the cam  102 , without adding mass. In another example, the component arrangement must be determined to fit within the package size constraints.   4. In step  353 , evaluate the results of step  351  and step  352 . If minimum requirements for the valve tip load  614  and dynamic stability at the selected engine speed are not met, iterate on the arrangement of components and perform the analyses in steps  351  and  352  again. When minimum requirements for the valve tip load  614  and dynamic stability at the selected engine speed are met, calculate deflection and stress for the rocker arm assembly  100 .   5. In step  354 , calculate stress and deflections.   6. In step  356 , evaluate deflection and stress. If minimum requirements for deflection and stress are not met, proceed to step  355 , and, and refine component design. When the design iteration is complete, return to step  353  and re-evaluate the valve tip load  614  and dynamic stability. When minimum requirements for the valve tip load  614  and dynamic stability at the selected engine speed are met, calculate deflection and stress in step  354 .   7. With reference to  FIG. 55 , when conditions of stress, deflection, and dynamic stability are met, the result is one possible design  357 . Analysis results can be plotted for possible design configurations on a graph of stiffness versus inertia. This graph provides a range of acceptable values as indicated by area  360 .  FIG. 57  shows three discrete acceptable designs. By extension, the acceptable inertia/stiffness area  360  also bounds the characteristics for individual major components  120 ,  622 , and torsion springs  134 ,  136 .       

     Now, with reference to  FIGS. 4, 52, 55 , a successful design, as described above, is reached if each of the major rocker arm assembly  100  components, including the outer arm  120 , the inner arm assembly  622 , and the torsion springs  134 ,  136 , collectively meet specific design criteria for inertia, stress, and deflection. A successful design produces unique characteristic data for each major component. 
     To illustrate, select three functioning DVVL rocker arm assemblies  100 , illustrated in  FIG. 57 , that meet a certain stiffness/inertia criteria. Each of these assemblies is comprised of three major components: the torsion springs  134 ,  136 , outer arm  120 , and inner arm assembly  622 . For this analysis, as illustrated in an exemplary illustration of  FIG. 58 , a range of possible inertia values for each major component can be described:
         Torsion spring set, design # 1 , inertia=A; torsion spring set, design # 2 , inertia=B; torsion spring set, design # 3 , inertia=C   Torsion spring set inertia range, calculated about the ball end plunger tip (also indicated with an X in  FIG. 59 ), is bounded by the extents defined in values A, B, and C.   Outer arm, design # 1 , inertia=D; outer arm, design # 2 , inertia=E; outer arm, design # 3 , inertia=F   Outer arm inertia range, calculated about the ball end plunger tip (also indicated with an X in  FIG. 59 ), is bounded by the extents defined in values D, E, and F   Inner arm assembly, design # 1 , inertia=X; inner arm assembly, design # 2 , inertia=Y; inner arm assembly, design # 3 , inertia=Z   Inner arm assembly inertia range, calculated about the ball end plunger tip (also indicated with an X in  FIG. 59 ), is bounded by the extents defined in values X, Y, and Z.       

     This range of component inertia values in turn produces a unique arrangement of major components (torsion springs, outer arm, and inner arm assembly). For example, in this design, the torsion springs will tend to be very close to the ball end plunger tip  611 . 
     As to  FIGS. 57-61 , calculation of inertia for individual components is closely tied to loading requirements in the assembly, because the desire to minimize inertia requires the optimization of mass distribution in the part to manage stress in key areas. For each of the three successful designs described above, a range of values for stiffness and mass distribution can be described.
         For outer arm  120  design # 1 , mass distribution can be plotted versus distance along the part, starting at end A, and proceeding to end B. In the same way, mass distribution values for outer arm  120  design # 2  and outer arm  120  design # 3  can be plotted.   The area between the two extreme mass distribution curves can be defined as a range of values characteristic to the outer arm  120  in this assembly.   For outer arm  120  design # 1 , stiffness distribution can be plotted versus distance along the part, starting at end A, and proceeding to end B. In the same way, stiffness values for outer arm  120  design # 2  and outer arm  120  design # 3  can be plotted.   The area between the two extreme stiffness distribution curves can be defined as a range of values characteristic to the outer arm  120  in this assembly.       

     Stiffness and mass distribution for the outer arm  120  along an axis related to its motion and orientation during operation, describe characteristic values, and by extension, characteristic shapes. 
     5 Design Verification 
     5.1 Latch Response 
     Latch response times for the exemplary DVVL system were validated with a latch response test stand  900  illustrated in  FIG. 62 , to ensure that the rocker arm assembly switched within the prescribed mechanical switching window explained previously, and illustrated in  FIG. 26 . Response times were recorded for oil temperatures ranging from 10° C. to 120° C. to effect a change in oil viscosity with temperature. 
     The latch response test stand  900  utilized production intent hardware including OCVs, DFHLAs, and DVVL switching rocker arms  100 . To simulate engine oil conditions, the oil temperature was controlled by an external heating and cooling system. Oil pressure was supplied by an external pump and controlled with a regulator. Oil temperature was measured in a control gallery between the OCV and DFHLA. The latch movement was measured with a displacement transducer  901 . 
     Latch response times were measured with a variety of production intent SRFFs. Tests were conducted with production intent 5w-20 motor oil. Response times were recorded when switching from low lift mode to high lift and high lift mode to low lift mode. 
       FIG. 21  details the latch response times when switching from low-lift mode to high-lift mode. The maximum response time at 20° C. was measured to be less than 10 milliseconds.  FIG. 22  details the mechanical response times when switching from high-lift mode to low lift mode. The maximum response time at 20° C. was measured to be less than 10 milliseconds. 
     Results from the switching studies show that the switching time for the latch is primarily a function of the oil temperature due to the change in viscosity of the oil. The slope of the latch response curve resembles viscosity to temperature relationships of motor oil. 
     The switching response results show that the latch movement is fast enough for mode switching in one camshaft revolution up to 3500 engine rpm. The response time begins to increase significantly as the temperature falls below 20° C. At temperatures of 10° C. and below, switching in one camshaft revolution is not possible without lowering the 3500 rpm switching requirement. 
     The SRFF was designed to be robust at high engine speeds for both high and low lift modes as shown in Table 1. The high lift mode can operate up to 7300 rpm with a “burst” speed requirement of 7500 rpm. A burst is defined as a short excursion to a higher engine speed. The SRFF is normally latched in high lift mode such that high lift mode is not dependent on oil temperature. The low lift operating mode is focused on fuel economy during part load operation up to 3500 rpm with an over speed requirement of 5000 rpm in addition to a burst speed to 7500 rpm. As tested, the system is able to hydraulically unlatch the SRFF for oil temperatures at 200 C or above. Testing was conducted down to 10° C. to ensure operation at 20° C. Durability results show that the design is robust across the entire operating range of engine speeds, lift modes and oil temperatures. 
     
       
         
           
               
               
               
               
             
               
                   
                 TABLE 1 
               
               
                   
                   
               
               
                   
                 Mode 
                 Engine Speed, rpm 
                 Oil Temperature 
               
               
                   
                   
               
             
            
               
                   
                 High Lift 
                 7300 
                 N/A 
               
               
                   
                   
                 7500 burst speed 
                   
               
               
                   
                 Low Lift 
                 3500 
                 20° C. and above 
               
               
                   
                 (Fuel Economy Mode) 
                 5000 overspeed 
                   
               
               
                   
                   
                 7500 burst speed 
               
               
                   
                   
               
            
           
         
       
     
     The design, development, and validation of a SRFF based DVVL system to achieve early intake valve closing was completed for a Type II valve train. This DVVL system improves fuel economy without jeopardizing performance by operating in two modes. Pumping loop losses are reduced in low lift mode by closing the intake valve early while performance is maintained in high lift mode by utilizing a standard intake valve profile. The system preserves common Type II intake and exhaust valve train geometries for use in an in-line four cylinder gasoline engine. Implementation cost is minimized by using common components and a standard chain drive system. Utilizing a Type II SRFF based system in this manner allows the application of this hardware to multiple engine families. 
     This DVVL system, installed on the intake of the valve train, met key performance targets for mode switching and dynamic stability in both high-lift and low-lift modes. Switching response times allowed mode switching within one cam revolution at oil temperatures above 20° C. and engine speeds up to 3500 rpm. Optimization of the SRFF stiffness and inertia, combined with an appropriate valve lift profile design allowed the system to be dynamically stable to 3500 rpm in low lift mode and 7300 rpm in high lift mode. The validation testing completed on production intent hardware shows that the DVVL system exceeds durability targets. Accelerated system aging tests were utilized to demonstrate durability beyond the life targets. 
     5.2 Durability 
     Passenger cars are required to meet an emissions useful life requirement of 150,000 miles. This study set a more stringent target of 200,000 miles to ensure that the product is robust well beyond the legislated requirement. 
     The valve train requirements for end of life testing are translated to the 200,000 mile target. This mileage target must be converted to valve actuation events to define the valve train durability requirements. In order to determine the number of valve events, the average vehicle and engine speeds over the vehicle lifetime must be assumed. For this example, an average vehicle speed of 40 miles per hour combined with an average engine speed of 2200 rpm was chosen for the passenger car application. The camshaft speed operates at half the engine speed and the valves are actuated once per camshaft revolution, resulting in a test requirement of 330 million valve events. Testing was conducted on both firing engines and non-firing fixtures. Rather than running a 5000 hour firing engine test, most testing and reported results focus on the use of the non-firing fixture illustrated in  FIG. 63  to conduct testing necessary to meet 330 million valve events. Results from firing and non-firing tests were compared, and the results corresponded well with regarding valve train wear results, providing credibility for non-firing fixture life testing. 
     5.2.1 Accelerated Aging 
     There was a need for conducting an accelerated test to show compliance over multiple engine lives prior to running engine tests. Hence, fixture testing was performed prior to firing tests. A higher speed test was designed to accelerate valve train wear such that it could be completed in less time. A test correlation was established such that doubling the average engine speed relative to the in-use speed yielded results in approximately one-quarter of the time and nearly equivalent valve train wear. As a result, valve train wear followed closely to the following equation: 
     
       
         
           
             
               VE 
               Accel 
             
             ≈ 
             
               
                 
                   VE 
                   
                     in 
                     ⁢ 
                     
                       - 
                     
                     ⁢ 
                     use 
                   
                 
                 ⁡ 
                 
                   ( 
                   
                     
                       RPM 
                       
                         avg 
                         ⁢ 
                         
                           - 
                         
                         ⁢ 
                         test 
                       
                     
                     
                       RPM 
                       
                         avg 
                         ⁢ 
                         
                           - 
                         
                         ⁢ 
                         in 
                         ⁢ 
                         
                             
                         
                         ⁢ 
                         use 
                       
                     
                   
                   ) 
                 
               
               2 
             
           
         
       
     
     Where VE Accel  are the valve events required during an accelerated aging test, VE in-use  are the valve events required during normal in-use testing, RPM avg-test  is the average engine speed for the accelerated test and RPM avg-in use  is the average engine speed for in-use testing. 
     A proprietary, high speed, durability test cycle was developed that had an average engine speed of approximately 5000 rpm. Each cycle had high speed durations in high lift mode of approximately 60 minutes followed by lower speed durations in low lift mode for approximately another 10 minutes. This cycle was repeated 430 times to achieve 72 million valve events at an accelerated wear rate that is equivalent to 330 million events at standard load levels. Standard valve train products containing needle and roller bearings have been used successfully in the automotive industry for years. This test cycle focused on the DLC coated slider pads where approximately 97% of the valve lift events were on the slider pads in high lift mode leaving 2 million cycles on the low lift roller bearing as shown in Table 2. These testing conditions consider one valve train life equivalent to 430 accelerated test cycles. Testing showed that the SRFF is durable through six engine useful lives with negligible wear and lash variation. 
     
       
         
           
               
             
               
                 TABLE 2 
               
             
            
               
                   
               
               
                 Durability Tests, Valve Events and Objectives 
               
            
           
           
               
               
               
               
            
               
                 Durability  
                 Duration 
                 Valve Events 
                   
               
            
           
           
               
               
               
               
               
            
               
                 Test 
                 (hours) 
                 total 
                 high lift 
                 Objective 
               
               
                   
               
            
           
           
               
               
               
               
               
            
               
                 Accelerated  
                 500 
                 72M 
                 97% 
                 Accelerated high speed  
               
               
                 System Aging 
                   
                   
                   
                 wear 
               
               
                 Switching 
                 500 
                 54M 
                 50% 
                 Latch and torsion spring  
               
               
                   
                   
                   
                   
                 wear 
               
               
                 Critical Shift 
                 800 
                 42M 
                 50% 
                 Lathe and bearing wear 
               
               
                 Idle 1 
                 1000 
                 27M 
                 100%  
                 Low lubrication 
               
               
                 Idle 2 
                 1000 
                 27M 
                  0% 
                 Low lubrication 
               
               
                 Cold Start 
                 1000 
                 27M 
                 100%  
                 Low lubrication 
               
               
                 Used Oil 
                 400 
                 56M 
                 ~99.5%     
                 Accelerated high speed  
               
               
                   
                   
                   
                   
                 wear 
               
               
                 Bearing 
                 140 
                 N/A 
                 N/A 
                 Bearing wear 
               
               
                 Torsion  
                 500 
                 25M 
                  0% 
                 Spring load loss 
               
               
                 Spring 
               
               
                   
               
            
           
         
       
     
     The accelerated system aging test was key to showing durability while many function-specific tests were also completed to show robustness over various operating states. 
     Table 2 includes the main durability tests combined with the objective for each test. The accelerated system aging test was described above showing approximately 500 hours or approximately 430 test cycles. A switching test was operated for approximately 500 hours to assess the latch and torsion spring wear. Likewise, a critical shift test was also performed to further age the parts during a harsh and abusive shift from the outer arm being partially latched such that it would slip to the low lift mode during the high lift event. A critical shift test was conducted to show robustness in the case of extreme conditions caused by improper vehicle maintenance. This critical shift testing was difficult to achieve and required precise oil pressure control in the test laboratory to partially latch the outer arm. This operation is not expected in-use as the oil control pressures are controlled outside of that window. Multiple idle tests combined with cold start operation were conducted to accelerate wear due to low oil lubrication. A used oil test was also conducted at high speed. Finally, bearing and torsion spring tests were conducted to ensure component durability. All tests met the engine useful lift requirement of 200,000 miles which is safely above the 150,000 mile passenger car useful life requirement. 
     All durability tests were conducted having specific levels of oil aeration. Most tests had oil aeration levels ranging between approximately 15% and 20% total gas content (TGC) which is typical for passenger car applications. This content varied with engine speed and the levels were quantified from idle to 7500 rpm engine speed. An excessive oil aeration test was also conducted having aeration levels of 26% TGC. These tests were conducted with SRFF&#39;s that met were tested for dynamics and switching performance tests. Details of the dynamics performance test are discussed in the results section. The oil aeration levels and extended levels were conducted to show product robustness. 
     5.2.2 Durability Test Apparatus 
     The durability test stand shown in  FIG. 63  consists of a prototype 2.5 L four cylinder engine driven by an electric motor with an external engine oil temperature control system  905 . Camshaft position is monitored by an Accu-coder  802 S external encoder  902  driven by the crankshaft Angular velocity of the crankshaft is measured with a digital magnetic speed sensor (model Honeywell 584)  904 . Oil pressure in both the control and hydraulic galleries is monitored using Kulite XTL piezoelectric pressure transducers. 
     5.2.3 Durability Test Apparatus Control 
     A control system for the fixture is configured to command engine speed, oil temperature and valve lift state as well as verify that the intended lift function is met. The performance of the valve train is evaluated by measuring valve displacement using non-intrusive Bently Nevada 3300XL proximity probes  906 . The proximity probes measure valve lift up to 2 mm at one-half camshaft degree resolution. This provides the information necessary to confirm the valve lift state and post process the data for closing velocity and bounce analysis. The test setup included a valve displacement trace that was recorded at idle speed to represent the baseline conditions of the SRFF and is used to determine the master profile  908  shown in  FIG. 64 . 
       FIG. 17  shows the system diagnostic window representing one switching cycle for diagnosing valve closing displacement. The OCV is commanded by the control system resulting in movement of the OCV armature as represented by the OCV current trace  881 . The pressure downstream of the OCV in the oil control gallery increases as shown by the pressure curve  880 ; thus, actuating the latch pin resulting in a change of state from high-lift to low-lift. 
       FIG. 64  shows the valve closing tolerance  909  in relation to the master profile  908  that was experimentally determined. The proximity probes  906  used were calibrated to measure the last 2 mm of lift, with the final 1.2 mm of travel shown on the vertical axis in  FIG. 64 . A camshaft angle tolerance of 2.5″ was established around the master profile  908  to allow for the variation in lift that results from valve train compression at high engine speeds to prevent false fault recording. A detection window was established to resolve whether or not the valve train system had the intended deflection. For example, a sharper than intended valve closing would result in an earlier camshaft angle closing resulting in valve bounce due to excessive velocity which is not desired. The detection window and tolerance around the master profile can detect these anomalies. 
     5.2.4 Durability Test Plan 
     A Design Failure Modes and Effects Analysis (DFMEA) was conducted to determine the SRFF failure modes. Likewise, mechanisms were determined at the system and subsystem levels. This information was used to develop and evaluate the durability of the SRFF to different operating conditions. The test types were separated into four categories as shown in  FIG. 65  that include: Performance Verification, Subsystem Testing, Extreme Limit Testing and Accelerated System Aging. 
     The hierarchy of key tests for durability are shown in  FIG. 65 . Performance Verification Testing benchmarks the performance of the SRFF to application requirements and is the first step in durability verification. Subsystem tests evaluate particular functions and wear interfaces over the product lifecycle. Extreme Limit Testing subjects the SRFF to the severe user in combination with operation limits. Finally, the Accelerated Aging test is a comprehensive test evaluating the SRFF holistically. The success of these tests demonstrates the durability of the SRFF. 
     Performance Verification 
     Fatigue &amp; Stiffness 
     The SRFF is placed under a cyclic load test to ensure fatigue life exceeds application loads by a significant design margin. Valve train performance is largely dependent on the stiffness of the system components. Rocker arm stiffness is measured to validate the design and ensure acceptable dynamic performance. 
     Valve train Dynamics 
     The Valve train Dynamics test description and performance is discussed in the results section. The test involved strain gaging the SRFF combined with measuring valve closing velocities. 
     Subsystem Testing 
     Switching Durability 
     The switching durability test evaluates the switching mechanism by cycling the SRFF between the latched, unlatched and back to the latched state a total of three million times ( FIGS. 24 and 25 ). The primary purpose of the test is the evaluation of the latching mechanism. Additional durability information is gained regarding the torsion springs due to 50% of the test cycle being in low lift. 
     Torsion Spring Durability and Fatigue 
     The torsion spring is an integral component of the switching roller finger follower. The torsion spring allows the outer arm to operate in lost motion while maintaining contact with the high lift camshaft lobe. The Torsion Spring Durability test is performed to evaluate the durability of the torsion springs at operational loads. The Torsion Spring Durability test is conducted with the torsion springs installed in the SRFF. The Torsion Spring Fatigue test evaluates the torsion spring fatigue life at elevated stress levels. Success is defined as torsion spring load loss of less than 15% at end-of-life. 
     Idle Speed Durability 
     The Idle Speed Durability test simulates a limit lubrication condition caused by low oil pressure and high oil temperature. The test is used to evaluate the slider pad and bearing, valve tip to valve pallet and ball socket to ball plunger wear. The lift-state is held constant throughout the test in either high or low lift. The total mechanical lash is measured at periodic inspection intervals and is the primary measure of wear. 
     Extreme Limit Testing 
     Overspeed 
     Switching rocker arm failure modes include loss of lift-state control. The SRFF is designed to operate at a maximum crankshaft speed of 3500 rpm in low lift mode. The SRFF includes design protection to these higher speeds in the case of unexpected malfunction resulting in low lift mode. Low lift fatigue life tests were performed at 5000 rpm. Engine Burst tests were performed to 7500 rpm for both high and low lift states. 
     Cold Start Durability 
     The Cold Start durability test evaluates the ability of the DLC to withstand 300 engine starting cycles from an initial temperature of −30° C. Typically, cold weather engine starting at these temperatures would involve an engine block heater. This extreme test was chosen to show robustness and was repeated 300 times on a motorized engine fixture. This test measures the ability of the DLC coating to withstand reduced lubrication as a result of low temperatures. 
     Critical Shift Durability 
     The SRFF is designed to switch on the base circle of the camshaft while the latch pin is not in contact with the outer arm. In the event of improper OCV timing or lower than required minimum control gallery oil pressure for full pin travel, the pin may still be moving at the start of the next lift event. The improper location of the latch pin may lead to a partial engagement between the latch pin and outer arm. In the event of a partial engagement between the outer arm and latch pin, the outer arm may slip off the latch pin resulting in an impact between the roller bearing and low lift camshaft lobe. The Critical Shift Durability is an abuse test that creates conditions to quantify robustness and is not expected in the life of the vehicle. The Critical Shift test subjects the SRFF to 5000 critical shift events. 
     Accelerated Bearing Endurance 
     The accelerated bearing endurance is a life test used to evaluate life of bearings that completed the critical shift test. The test is used to determine whether the effects of critical shift testing will shorten the life of the roller bearing. The test is operated at increased radial loads to reduce the time to completion. New bearings were tested simultaneously to benchmark the performance and wear of the bearings subjected to critical shift testing. Vibration measurements were taken throughout the test and were analyzed to detect inception of bearing damage. 
     Used Oil Testing 
     The Accelerated System Aging test and Idle Speed Durability test profiles were performed with used oil that had a 20/19/16 ISO rating. This oil was taken from engines at the oil change interval. 
     Accelerated System Aging 
     The Accelerated System Aging test is intended to evaluate the overall durability of the rocker arm including the sliding interface between the camshaft and SRFF, latching mechanism and the low lift bearing. The mechanical lash was measured at periodic inspection intervals and is the primary measure of wear.  FIG. 66  shows the test protocol in evaluating the SRFF over an Accelerated System Aging test cycle. The mechanical lash measurements and FTIR measurements allow investigation of the overall health of the SRFF and the DLC coating respectively. Finally, the part is subjected to a teardown process in an effort to understand the source of any change in mechanical lash from the start of test. 
       FIG. 67  is a pie chart showing the relative testing time for the SRFF durability testing which included approximately 15,700 total hours. The Accelerated System Aging test offered the most information per test hour due to the acceleration factor and combined load to the SRFF within one test leading to the 37% allotment of total testing time. The Idle Speed Durability (Low Speed, Low Lift and Low Speed, High Lift) tests accounted for 29% of total testing time due to the long duration of each test. Switching Durability was tested to multiple lives and constituted 9% of total test time. Critical Shift Durability and Cold Start Durability testing required significant time due to the difficulty in achieving critical shifts and thermal cycling time required for the Cold Start Durability. The data is quantified in terms of the total time required to conduct these modes as opposed to just the critical shift and cold starting time itself. The remainder of the subsystem and extreme limit tests required 11% of the total test time. 
     Valvetrain Dynamics 
     Valve train dynamic behavior determines the performance and durability of an engine. Dynamic performance was determined by evaluating the closing velocity and bounce of the valve as it returns to the valve seat. Strain gaging provides information about the loading of the system over the engine speed envelope with respect to camshaft angle. Strain gages are applied to the inner and outer arms at locations of uniform stress.  FIG. 68  shows a strain gage attached to the SRFF. The outer and inner arms were instrumented to measure strain for the purpose of verifying the amount of load on the SRFF. 
     A Valve train Dynamics test was conducted to evaluate the performance capabilities of the valve train. The test was performed at nominal and limit total mechanical lash values. The nominal case is presented. A speed sweep from 1000 to 7500 rpm was performed, recording 30 valve events per engine speed. Post processing of the dynamics data allows calculation of valve closing velocity and valve bounce. The attached strain gages on the inner and outer arms of the SRFF indicate sufficient loading of the rocker arm at all engine speeds to prevent separation between valve train components or “pump-up” of the HLA. Pump-up occurs when the HLA compensates for valve bounce or valve train deflection causing the valve to remain open on the camshaft base circle. The minimum, maximum and mean closing velocities are shown to understand the distribution over the engine speed range. The high lift closing velocities are presented in  FIG. 67 . The closing velocities for high lift meet the design targets. The span of values varies by approximately 250 mm/s between the minimum and maximum at 7500 rpm while safely staying within the target. 
       FIG. 69  shows the closing velocity of the low lift camshaft profile. Normal operation occurs up to 3500 rpm where the closing velocities remain below 200 mm/s, which is safely within the design margin for low lift. The system was designed to an over-speed condition of 5000 rpm in low lift mode where the maximum closing velocity is below the limit. Valve closing velocity design targets are met for both high and low lift modes. 
     Critical Shift 
     The Critical Shift test is performed by holding the latch pin at the critical point of engagement with the outer arm as shown in  FIG. 27 . The latch is partially engaged on the outer arm which presents the opportunity for the outer arm to disengage from the latch pin resulting in a momentary loss of control of the rocker arm. The bearing of the inner arm is impacted against the low lift camshaft lobe. The SRFF is tested to a quantity that far exceeds the number of critical shifts that are anticipated in a vehicle to show lifetime SRFF robustness. The Critical Shift test evaluates the latching mechanism for wear during latch disengagement as well as the bearing durability from the impact that occurs during a critical shift. 
     The Critical Shift test was performed using a motorized engine similar to that shown in  FIG. 63 . The lash adjuster control gallery was regulated about the critical pressure. The engine is operated at a constant speed and the pressure is varied around the critical pressure to accommodate for system hysteresis. A Critical Shift is defined as a valve drop of greater than 1.0 mm. The valve drop height distribution of a typical SRFF is shown in  FIG. 70 . It should be noted that over 1000 Critical Shifts occurred at less than 1.0 mm which are tabulated but not counted towards test completion.  FIG. 71  displays the distribution of critical shifts with respect to camshaft angle. The largest accumulation occurs immediately beyond peak lift with the remainder approximately evenly distributed. 
     The latching mechanism and bearing are monitored for wear throughout the test. The typical wear of the outer arm ( FIG. 73 ) is compared to a new part ( FIG. 72 ). Upon completion of the required critical shifts, the rocker arm is checked for proper operation and the test concluded. The edge wear shown did not have a significant effect on the latching function and the total mechanical lash as the majority of the latch shelf displayed negligible wear. 
     Subsystems 
     The subsystem tests evaluate particular functions and wear interfaces of the SRFF rocker arm. Switching Durability evaluates the latching mechanism for function and wear over the expected life of the SRFF Similarly, Idle Speed Durability subjects the bearing and slider pad to a worst case condition including both low lubrication and an oil temperature of 130° C. The Torsion Spring Durability Test was accomplished by subjecting the torsion springs to approximately 25 million cycles. Torsion spring loads are measured throughout the test to measure degradation. Further confidence was gained by extending the test to 100 million cycles while not exceeding the maximum design load loss of 15%.  FIG. 74  displays the torsion spring loads on the outer arm at start and end of test. Following 100 million cycles, there was a small load loss on the order of 5% to 10% which is below the 15% acceptable target and shows sufficient loading of the outer arm to four engine lives. 
     Accelerated System Aging 
     The Accelerated System Aging test is the comprehensive durability test used as the benchmark of sustained performance. The test represents the cumulative damage of the severe end-user. The test cycle averages approximately 5000 rpm with constant speed and acceleration profiles. The time per cycle is broken up as follows: 28% steady state, 15% low lift and cycling between high and low lift with the remainder under acceleration conditions. The results of testing show that the lash change in one-life of testing accounts for 21% of the available wear specification of the rocker arm. Accelerated System Aging test, consisting of 8 SRFF&#39;s, was extended out past the standard life to determine wear out modes of the SRFF. Total mechanical lash measurements were recorded every 100 test cycles once past the standard duration. 
     The results of the accelerated system aging measurements are presented in  FIG. 75  showing that the wear specification was exceeded at 3.6 lives. The test was continued and achieved six lives without failure. Extending the test to multiple lives displayed a linear change in mechanical lash once past an initial break in period. The dynamic behavior of the system degraded due to the increased total mechanical lash; nonetheless, functional performance remained intact at six engine lives. 
     5.2.5 Durability Test Results 
     Each of the tests discussed in the test plan were performed and a summary of the results are presented. The results of Valve train Dynamics, Critical Shift Durability, Torsion Spring Durability and finally the Accelerated System Aging test are shown. 
     The SRFF was subjected to accelerated aging tests combined with function-specific tests to demonstrate robustness and is summarized in Table 3. 
     
       
         
           
               
             
               
                 TABLE 3 
               
             
            
               
                   
               
               
                 Durability Summary 
               
            
           
           
               
               
               
               
            
               
                   
                   
                   
                 Valve Events 
               
            
           
           
               
               
               
               
               
            
               
                 Durability Test 
                 Lifetimes 
                 Cycles 
                 total 
                 # tests 
               
               
                   
               
            
           
           
               
               
               
               
               
            
               
                 Accelerated System Aging 
                   6 
                   
                   
                   
               
               
                 Switching 
                   1 (used oil) 
                   
                   
                   
               
               
                 Torsion Spring 
                   3 
                   
                   
                   
               
               
                 Critical Shift 
                   4 
                   
                   
                   
               
               
                 Cold Start 
                 &gt;1 
                   
                   
                   
               
               
                 Overspeed 
                 &gt;1 
                   
                   
                   
               
               
                 (5000 rpm in low lift) 
                   
                   
                   
                   
               
               
                 Overspeed 
                 &gt;1 
                   
                   
                   
               
               
                 (7500 rpm in high lift) 
                   
                   
                   
                   
               
               
                 Bearing 
                   
                   
                 100M  
                 1 
               
               
                 Idle low lift 
                   
                   
                 27M 
                 2 
               
               
                 Idle high lift 
                 &gt;1 
                   
                 27M 
                 2 
               
               
                   
                 &gt;1 (dirty oil) 
                   
                 27M 
                 1 
               
               
                   
               
               
                 Legend: 
               
               
                 1 engine lifetime = 200,000 miles (safe margin over the 150,000 mile requirement) 
               
            
           
         
       
     
     Durability was assessed in terms of engine lives totaling an equivalent 200,000 miles which provides substantial margin over the mandated 150,000 mile requirement. The goal of the project was to demonstrate that all tests show at least one engine life. The main durability test was the accelerated system aging test that exhibited durability to at least six engine lives or 1.2 million miles. This test was also conducted with used oil showing robustness to one engine life. A key operating mode is switching operation between high and low lift. The switching durability test exhibited at least three engine lives or 600,000 miles. Likewise, the torsion spring was robust to at least four engine lives or 800,000 miles. The remaining tests were shown to at least one engine life for critical shifts, over speed, cold start, bearing robustness and idle conditions. The DLC coating was robust to all conditions showing polishing with minimal wear, as shown in  FIG. 76 . As a result, the SRFF was tested extensively showing robustness well beyond a 200,000 mile useful life. 
     5.2.6 Durability Test Conclusions 
     The DVVL system including the SRFF, DFHLA and OCV was shown to be robust to at least 200,000 miles which is a safe margin beyond the 150,000 mile mandated requirement. The durability testing showed accelerated system aging to at least six engine lives or 1.2 million miles. This SRFF was also shown to be robust to used oil as well as aerated oil. The switching function of the SRFF was shown robust to at least three engine lives or 600,000 miles. All sub-system tests show that the SRFF was robust beyond one engine life of 200,000 miles. 
     Critical shift tests demonstrated robustness to 5000 events or at least one engine life. This condition occurs at oil pressure conditions outside of the normal operating range and causes a harsh event as the outer arm slips off the latch such that the SRFF transitions to the inner arm. Even though the condition is harsh, the SRFF was shown robust to this type of condition. It is unlikely that this event will occur in serial production. Testing results show that the SRFF is robust to this condition in the case that a critical shift occurs. 
     The SRFF was proven robust for passenger car application having engine speeds up to 7300 rpm and having burst speed conditions to 7500 rpm. The firing engine tests had consistent wear patterns to the non-firing engine tests described in this paper. The DLC coating on the outer arm slider pads was shown to be robust across all operating conditions. As a result, the SRFF design is appropriate for four cylinder passenger car applications for the purpose of improving fuel economy via reduced engine pumping losses at part load engine operation. This technology could be extended to other applications including six cylinder engines. The SRFF was shown to be robust in many cases that far exceeded automotive requirements. Diesel applications could be considered with additional development to address increased engine loads, oil contamination and lifetime requirements. 
     5.3 Slider Pad/DLC Coating Wear 
     5.3.1 Wear Test Plan 
     This section describes the test plan utilized to investigate the wear characteristics and durability of the DLC coating on the outer arm slider pad. The goal was to establish relationships between design specifications and process parameters and how each affected the durability of the sliding pad interface. Three key elements in this sliding interface are: the camshaft lobe, the slider pad, and the valve train loads. Each element has factors which needed to be included in the test plan to determine the effect on the durability of the DLC coating. Detailed descriptions for each component follow: 
     Camshaft—The width of the high lift camshaft lobes were specified to ensure the slider pad stayed within the camshaft lobe during engine operation. This includes axial positional changes resulting from thermal growth or dimensional variation due to manufacturing. As a result, the full width of the slider pad could be in contact with the camshaft lobe without risk of the camshaft lobe becoming offset to the slider pad. The shape of the lobe (profile) pertaining to the valve lift characteristics had also been established in the development of the camshaft and SRFF. This left two factors which needed to be understood relative to the durability of the DLC coating; the first was lobe material and the second was the surface finish of the camshaft lobe. The test plan included cast iron and steel camshaft lobes tested with different surface conditions on the lobe. The first included the camshafts lobes as prepared by a grinding operation (as-ground). The second was after a polishing operation improved the surface finish condition of the lobes (polished). 
     Slider Pad—The slider pad profile was designed to specific requirements for valve lift and valve train dynamics  FIG. 77  is a graphic representation of the contact relationship between the slider pads on the SRFF and the contacting high lift lobe pair. Due to expected manufacturing variations, there is an angular alignment relationship in this contacting surface which is shown in the  FIG. 77  in exaggerated scale. The crowned surface reduces the risk of edge loading the slider pads considering various alignment conditions. However, the crowned surface adds manufacturing complexity, so the effect of crown on the coated interface performance was added to the test plan to determine its necessity. 
     The  FIG. 77  shows the crown option on the camshaft surface as that was the chosen method. Hertzian stress calculations based on expected loads and crown variations were used for guidance in the test plan. A tolerance for the alignment between the two pads (included angle) needed to be specified in conjunction with the expected crown variation. The desired output of the testing was a practical understanding of how varying degrees of slider pad alignment affected the DLC coating. Stress calculations were used to provide a target value of misalignment of 0.2 degrees. These calculations served only as a reference point. The test plan incorporated three values for included angles between the slider pads: &lt;0.05 degrees, 0.2 degrees and 0.4 degrees. Parts with included angles below 0.05 degrees are considered flat and parts with 0.4 degrees represent a doubling of the calculated reference point. 
     The second factor on the slider pads which required evaluation was the surface finish of the slider pads before DLC coating. The processing steps of the slider pad included a grinding operation which formed the profile of the slider pad and a polishing step to prepare the surface for the DLC coating. Each step influenced the final surface finish of the slider pad before DLC coating was applied. The test plan incorporated the contribution of each step and provided results to establish an in-process specification for grinding and a final specification for surface finish after the polishing step. The test plan incorporated the surface finish as ground and after polish. 
     Valve train load—The last element was the loading of the slider pad by operation of the valve train. Calculations provided a means to transform the valve train loads into stress levels. The durability of both the camshaft lobe and the DLC coating was based on the levels of stress each could withstand before failure. The camshaft lobe material should be specified in the range of 800-1000 MPa (kinematic contact stress). This range was considered the nominal design stress. In order to accelerate testing, the levels of stress in the test plan were set at 900-1000 MPa and 1125-1250 MPa. These values represent the top half of the nominal design stress and 125% of the design stress respectively. 
     The test plan incorporated six factors to investigate the durability of the DLC coating on the slider pads: (1) the camshaft lobe material, (2) the form of the camshaft lobe, (3) the surface conditions of the camshaft lobe, (4) the angular alignment of the slider pad to the camshaft lobe, {S} the surface finish of the slider pad and (6) the stress applied to the coated slider pad by opening the valve. A summary of the elements and factors outlined in this section is shown in Table 1. 
     
       
         
           
               
             
               
                 TABLE 1 
               
             
            
               
                   
               
               
                 Test Plan Elements and Factors 
               
            
           
           
               
               
            
               
                 Element 
                 Factor 
               
               
                   
               
               
                 Camshaft 
                 Material: Cast Iron, steel 
               
               
                   
                 Surface Finish: as ground, polished 
               
               
                   
                 Lobe Form: Flat, Crowned 
               
               
                 Slider Pad 
                 Angular Alignment: &lt;0.05, 0.2, 0.4 degrees 
               
               
                   
                 Surface Finish: as ground, polished 
               
               
                 Valvetrain Load 
                 Stress Level: Max Design, 125% Max Design 
               
               
                   
               
            
           
         
       
     
     5.3.2 Component Wear Test Results 
     The goal of testing was to determine relative contribution each of the factors had on the durability of the slider pad DLC coating. The majority of the test configurations included a minimum of two factors from the test plan. The slider pads  752  were attached to a support rocker  753  on a test coupon  751  shown in  FIG. 78 . All the configurations were tested at the two stress levels to allow for a relative comparison of each of the factors. Inspection intervals ranged from 20-50 hours at the start of testing and increased to 300-500 hour intervals as results took longer to observe. Testing was suspended when the coupons exhibited loss of the DLC coating or there was a significant change in the surface of the camshaft lobe. The testing was conducted at stress levels higher than the application required hastening the effects of the factors. As a result, the engine life assessment described is a conservative estimate and was used to demonstrate the relative effect of the tested factors. Samples completing one life on the test stand were described as adequate. Samples exceeding three lives without DLC loss were considered excellent. The test results were separated into two sections to facilitate discussion. The first section discusses results from the cast iron camshafts and the second examines results from the steel camshafts. 
     Test Results for Cast Iron Camshafts 
     The first tests utilized cast iron camshaft lobes and compared slider pad surface finish and two angular alignment configurations. The results are shown in Table 2 below. This table summarizes the combinations of slider pad included angle and surface conditions tested with the cast iron camshafts. Each combination was tested at the max: design and 125% max design load condition. The values listed represent the number of engine lives each combination achieved during testing. 
     
       
         
           
               
             
               
                 TABLE 2 
               
               
                   
               
               
                 Cast Iron Test Matrix and Results 
               
               
                 Cast Iron Camshaft 
               
               
                   
               
             
            
               
                   
               
            
           
           
               
               
               
            
               
                 Lobe Surface Finish 
                 Ground 
                   
               
               
                 Lobe Profile 
                 Flat 
                   
               
            
           
           
               
               
               
               
               
               
            
               
                 Slider Pad 
                 0.2 deg. 
                 Ground 
                 0.1 
                 0.1 
                 Engine 
               
               
                 Configuration 
                   
                 Polished 
                 0.5 
                 0.3 
                 Lives 
               
               
                   
                 Flat 
                 Ground 
                 0.3 
                 0.2 
                   
               
               
                   
                   
                 Polished 
                 0.75 
                 0.4 
                   
               
               
                   
                 Included  
                 Surface 
                 Max  
                 125% Max 
                   
               
               
                   
                 Angle 
                 Preparation 
                 Design 
                 Design 
                   
               
            
           
           
               
               
               
               
               
            
               
                   
                   
                   
                 Valvetrain Load 
                   
               
               
                   
               
            
           
         
       
     
     The camshafts from the tests all developed spalling which resulted in the termination of the tests. The majority developed spalling before half an engine life. The spalling was more severe on the higher load parts but also present on the max design load parts. Analysis revealed both loads exceeded the capacity of the camshaft. Cast iron camshaft lobes are commonly utilized in applications with rolling elements containing similar load levels; however, in this sliding interface, the material was not a suitable choice. 
     The inspection intervals were frequent enough to study the effect the surface finish had on the durability of the coating. The coupons with the as-ground surface finish suffered DLC coating loss very early in the testing. The coupon shown in  FIG. 79A  illustrates a typical sample of the DLC coating loss early in the test. 
     Scanning electron microscope (SEM) analysis revealed the fractured nature of the DLC coating. The metal surface below the DLC coating did not offer sufficient support to the coating. The coating is significantly harder than the metal to which it is bonded; thus, if the base metal significantly deforms the DLC may fracture as a result. The coupons that were polished before coating performed well until the camshaft lobes started to spall. The best result for the cast iron camshafts was 0.75 lives with the combination of the flat, polished coupons at the max design load. 
     Test Results for Steel Camshafts 
     The next set of tests incorporated the steel lobe camshafts. A summary of the test combinations and results is listed in Table 3. The camshaft lobes were tested with four different configurations: (1) surface finish as ground with flat lobes, (2) surface finish as ground with crowned lobes, (3) polished with minimum crowned lobes and (4) polished with nominal crown on the lobes. The slider pads on the coupons were polished before DLC coating and tested at three angles: (1) flat (less than 0.05 degrees of included angle), (2) 0.2 degrees of included angle and (3) 0.4 degrees of included angle. The loads for all the camshafts were set at max design or 125% of the max design level. 
     
       
         
           
               
             
               
                 TABLE 3 
               
             
            
               
                   
               
               
                 Steel Camshaft Test Matrix and Results 
               
            
           
           
               
               
               
               
            
               
                 Lobe Surface Finish 
                 Ground 
                 Polished 
                   
               
            
           
           
               
               
            
               
                 Steel Camshaft 
                   
               
            
           
           
               
               
               
            
               
                   
                 Crown 
                   
               
            
           
           
               
               
               
               
               
            
               
                 Lobe Profile 
                 Flat 
                 Minimum 
                 Nominal 
               
               
                   
               
            
           
           
               
               
               
               
               
               
               
               
               
               
            
               
                 Slider Pad 
                 0.4 deg. 
                 Polished 
                 0.1 
                 0.75 
                 1.5 
                 2.3 
                 2.9 
                 2.6 
                 Engine 
               
               
                 Configuration 
                 0.2 deg. 
                 Polished 
                 1.6 
                 — 
                 3.3 
                 2.8 
                 3.1 
                 3 
                 Lives 
               
               
                   
                 Flat 
                 Polished 
                 — 
                 1.8  
                 2.6 
                 2.2 
                 3.3 
                 3 
               
               
                   
                 Included 
                 Surface 
                 Max 
                 125% 
                 Max 
                 125% 
                 Max 
                 125% 
               
               
                   
                 Angle 
                 Preparation 
                 Design 
                 Max 
                 Design 
                 Max 
                 Design 
                 Max 
               
               
                   
                   
                   
                   
                 Design 
                   
                 Design 
                   
                 Design 
               
            
           
           
               
               
               
            
               
                   
                 Valve train Load 
               
               
                   
                   
               
            
           
         
       
     
     The test samples which incorporated as-ground flat steel camshaft lobes and 0.4 degree included angle coupons at the 125% design load levels did not exceed one life. The samples tested at the maximum design stress lasted one life but exhibited the same effects on the coating. The 0.2 degree and flat samples performed better but did not exceed two lives. 
     This test was followed with ground, flat, steel camshaft lobes and coupons with 0.2 degree included angle and flat coupons. The time required before observing coating loss on the 0.2 degree samples was 1.6 lives. The flat coupons ran slightly longer achieving 1.8 lives. The pattern of DLC loss on the flat samples was non-uniform with the greatest losses on the outside of the contact patch. The loss of coating on the outside of the contact patches indicated the stress experienced by the slider pad was not uniform across its width. This phenomenon is known as “edge effect”. The solution for reducing the stress at the edges of two aligned elements is to add a crown profile to one of the elements. The application utilizing the SRFF has the crowned profile added to the camshaft. 
     The next set of tests incorporated the minimum value of crown combined with 0.4, 0.2 degree and flat polished slider pads. This set of tests demonstrated the positive consequence of adding crown to the camshaft. The improvement in the 125% max load was from 0.75 to 1.3 lives for the 0.4 degree samples. The flat parts exhibited a smaller improvement from 1.8 to 2.2 lives for the same load. 
     The last set of tests included all three angles of coupons with polished steel camshaft lobes machined with nominal crown values. The most notable difference in these results is the interaction between camshaft crown and the angular alignment of the slider pads to the camshaft lobe. The flat and 0.2 degree samples exceeded three lives at both load levels. The 0.4 degree samples did not exceed two lives.  FIG. 79B  shows a typical example of one of the coupons tested at the max design load with 0.2 degrees of included angle. 
     These results demonstrated the following: (1) the nominal value of camshaft crown was effective in mitigating slider pad angular alignment up to 0.2 degrees to flat; (2) the mitigation was effective at max design loads and 125% max design loads of the intended application and, (3) polishing the camshaft lobes contributes to the durability of the DLC coating when combined with slider pad polish and camshaft lobe crown. 
     Each test result helped to develop a better understanding of the effect stress had on the durability of the DLC coating. The results are plotted in  FIG. 80 . 
     The early tests utilizing cast iron camshaft lobes did not exceed half an engine life in a sliding interface at the design loads. The next improvement came in the form of identifying ‘edge effect’. The addition of crown to the polished camshaft lobes combined with a better understanding of allowable angular alignment, improved the coating durability to over three lives. The outcome is a demonstrated design margin between the observed test results and the maximum design stress for the application at each estimated engine life. 
     The effect surface finish has on DLC durability is most pronounced in the transition from coated samples as-ground to coated coupons as-polished. Slider pads tested as-ground and coated did not exceed one third engine life as shown in  FIG. 81 . Improvements in the surface finish of the slider pad provided greater load carrying capability of the substrate below the coating and improved overall durability of the coated slider pad. 
     The results from the cast iron and steel camshaft testing provided the following: (1) a specification for angular alignment of the slider pads to the camshaft, (2) clear evidence that the angular alignment specification was compatible with the camshaft lobe crown specification, (3) the DLC coating will remain intact within the design specifications for camshaft lobe crown and slider pad alignment beyond the maximum design load, (4) a polishing operation is required after the grinding of the slider pad, (5) an in-process specification for the grinding operation, (6) a specification for surface finish of the slider pads prior to coating and (7) a polish operation on the steel camshaft lobes contributes to the durability of the DLC coating on the slider pad. 
     5.4 Slider Pad Manufacturing Development 
     5.4.1 Slider Pad Manufacturing Development Description 
     The outer arm utilizes a machined casting. The prototype parts, machined from billet stock, had established targets for angular variation of the slider pads and the surface finish before coating. The development of the production grinding and polishing processes took place concurrently to the testing, and is illustrated in  FIG. 82 . The test results provided feedback and guidance in the development of the manufacturing process of the outer arm slider pad. Parameters in the process were adjusted based on the results of the testing and new samples machined were subsequently evaluated on the test fixture. 
     This section describes the evolution of the manufacturing process for the slider pad from the coupon to the outer arm of the SRFF. 
     The first step to develop the production grinding process was to evaluate different machines. A trial run was conducted on three different grinding machines. Each machine utilized the same vitrified cubic boron nitride (CBN) wheel and dresser. The CBN wheel was chosen as it offers (1) improved part to part consistency, (2) improved accuracy in applications requiring tight tolerances and (3) improved efficiency by producing more pieces between dress cycles compared to aluminum oxide. Each machine ground a population of coupons using the same feed rate and removing the same amount of material in each pass. A fixture was provided allowing the sequential grinding of coupons. The trial was conducted on coupons because the samples were readily polished and tested on the wear rig. This method provided an impartial means to evaluate the grinders by holding parameters like the fixture, grinding wheel and dresser as constants. 
     Measurements were taken after each set of samples were collected. Angular measurements of the slider pads were obtained using a Leitz PMM 654 coordinate measuring machine (CMM). Surface finish measurements were taken on a Mahr LD 120 profilometer.  FIG. 83  shows the results of the slider pad angle control relative to the grinder equipment. The results above the line are where a noticeable degradation of coating performance occurred. The target region indicates that the parts tested to this included angle show no difference in life testing. Two of the grinders failed to meet the targets for included angle of the slider pad on the coupons. The third did very well by comparison. The test results from the wear rig confirmed the sliding interface was sensitive to included angles above this target. The combination of the grinder trials and the testing discussed in the previous section helped in the selection of manufacturing equipment. 
       FIG. 84  summarizes the surface finish measurements of the same coupons as the included angle data shown in  FIG. 83 . The surface finish specification for the slider pads was established as a result of these test results. Surface finish values above the limit line shown have reduced durability. 
     The same two grinders (A and B) also failed to meet the target for surface finish. The target for surface finish was established based on the net change of surface finish in the polishing process for a given population of parts. Coupons that started out as outliers from the grinding process remained outliers after the polishing process; therefore, controlling surface finish at the grinding operation was important to be able to produce a slider pad after polish that meets the final surface finish prior to coating. 
     The measurements were reviewed for each machine. Grinders A and B both had variation in the form of each pad in the angular measurements. The results implied the grinding wheel moved vertically as it ground the slider pads. Vertical wheel movement in this kind of grinder is related to the overall stiffness of the machine. Machine stiffness also can affect surface finish of the part being ground. Grinding the slider pads of the outer arm to the specifications validated by the test fixture required the stiffness identified in Grinder C. 
     The lessons learned grinding coupons were applied to development of a fixture for grinding the outer arm for the SRFF. However the outer arm offered a significantly different set of challenges. The outer arm is designed to be stiff in the direction it is actuated by the camshaft lobes. The outer arm is not as stiff in the direction of the slider pad width. 
     The grinding fixture needed to (1) damp each slider pad without bias, (2) support each slider pad rigidly to resist the forces applied by grinding and (3) repeat this procedure reliably in high volume production. 
     The development of the outer arm fixture started with a manual clamping style block. Each revision of the fixture attempted to remove bias from the damping mechanism and reduce the variation of the ground surface.  FIG. 85  illustrates the results through design evolution of the fixture that holds the outer arm during the slider pad grinding operation. 
     The development completed by the test plan set boundaries for key SRFF outer arm slider pad specifications for surface finish parameters and form tolerance in terms of included angle. The influence of grind operation surface finish to resulting final surface finish after polishing was studied and used to establish specifications for the intermediate process standards. These parameters were used to establish equipment and part fixture development that assure the coating performance will be maintained in high volume production. 
     5.4.2 Slider Pad Manufacturing Development 
     Conclusions 
     The DLC coating on the SRFF slider pads that was configured in a DVVL system including DFHLA and OCV components was shown to be robust and durable well beyond the passenger car lifetime requirement. Although DLC coating has been used in multiple industries, it had limited production for the automotive valve train market. The work identified and quantified the effect of the surface finish prior to the DLC application, DLC stress level and the process to manufacture the slider pads. This technology was shown to be appropriate and ready for the serial production of a SRFF slider pad. 
     The surface finish was critical to maintaining DLC coating on the slider pads throughout lifetime tests. Testing results showed that early failures occurred when the surface finish was too rough. The paper highlighted a regime of surface finish levels that far exceeded lifetime testing requirements for the Ole This recipe maintained the DLC intact on top of the chrome nitride base layer such that the base metal of the SRFF was not exposed to contacting the camshaft lobe material. 
     The stress level on the DLC slider pad was also identified and proven. The testing highlighted the need for angle control for the edges of the slider pad. It was shown that a crown added to the camshaft lobe adds substantial robustness to edge loading effects due to manufacturing tolerances. Specifications set for the angle control exhibited testing results that exceeded lifetime durability requirements. 
     The camshaft lobe material was also found to be an important factor in the sliding interface. The package requirements for the SRFF based DVVL system necessitated a robust solution capable of sliding contact stresses up to 1000 MPa. The solution at these stress levels, a high quality steel material, was needed to avoid camshaft lobe spalling that would compromise the life of the sliding interface. The final system with the steel camshaft material, crowned and polished was found to exceed lifetime durability requirements. 
     The process to produce the slider pad and DLC in a high volume manufacturing process was discussed. Key manufacturing development focused on grinding equipment selection in combination with the grinder abrasive wheel and the fixture that holds the SRFF outer arm for the production slider pad grinding process. The manufacturing processes selected show robustness to meeting the specifications for assuring a durable sliding interface for the lifetime of the engine. 
     The DLC coating on the slider pads was shown to exceed lifetime requirements which are consistent with the system DVVL results. The DLC coating on the outer arm slider pads was shown to be robust across all operating conditions. As a result, the SRFF design is appropriate for four cylinder passenger car applications for the purpose of improving fuel economy via reduced engine pumping losses at part load engine operation. The DLC coated sliding interface for a DVVL was shown to be durable and enables VVA technologies to be utilized in a variety of engine valve train applications. 
     II. Single-Lobe Cylinder Deactivation System (CDA-1L) System Embodiment Description 
     1. CDA-1L System Overview 
     CDA-1L ( FIG. 88 ) is a compact cam-driven single-lobe cylinder deactivation (CDA-1L) switching rocker arm  1100  installed on a piston-driven internal combustion engine, and actuated with the combination of dual-feed hydraulic lash adjusters (DFHLA)  110  and oil control valves (OCV)  822 . 
     Now, in reference to  FIGS. 11, 88, 99, and 100 , the CDA-1L layout includes four main components: Oil control valve (OCV)  822 , dual feed hydraulic lash adjuster (DFHLA), CDA-1L switching rocker arm assembly (also referred to SRFF-1L)  1100 ; single-lobe cam  1320 . The default configuration is in the normal-lift (latched) position where the inner arm  1108  and outer arm  1102  of the CDA-1L rocker arm  1100  are locked together, causing the engine valve to open and allowing the cylinder to operate as it would in a standard valvetrain. The DFHLA  110  has two oil ports. The lower oil port  512  provides lash compensation and is fed engine oil similar to a standard HLA. The upper oil port  506 , referred as the switching pressure port, provides the conduit between controlled oil pressure from the OCV  822  and the latch  1202  in the SRFF-1L. As noted, when the latch is engaged, the inner arm  1108  and outer arm  1102  in the SRFF-1L  1110  operate together like a standard rocker arm to open the engine valve. In the no-lift (unlatched) position, the inner arm  1108  and outer arm  1102  can move independently to enable cylinder deactivation. 
     As shown in  FIGS. 88 and 99 , a pair of lost motion torsion springs  1124  are incorporated to bias the position of the inner arm  1108  so that it always maintains continuous contact with the camshaft lobe  1320 . The lost motion torsion springs  1124  require a higher preload than designs that use multiple lobes to facilitate continuous contact between the camshaft lobe  1320  and the inner arm roller bearing  1116 . 
       FIG. 89  shows a detailed view of the inner arm  1108  and outer arm  1102  in the SRFF-1L  1100  along with the latch  1202  mechanism and roller bearing  1116 . The functionality of the SRFF-1L  1100  design maintains similar packaging and reduces the complexity of the camshaft  1300  compared to configurations with more than one lobe, for example, separate no-lift lobes for each SRFF position can be eliminated. 
     As illustrated in  FIG. 91 , a complete CDA system  1400  for one engine cylinder includes one OCV  822 , two SRFF-1L rocker arms  1100  for the exhaust, two SRFF-1L rocker arms  1100  for the intake, one DFHLA  110  for each SRFF-1L  1100  and a single-lobe camshaft  1300  that drives each SRFF-1L  1100 . Additionally, the CDA  1400  system is designed such that the SRFF-1L  1100  and DFHLA  110  are identical for both the intake and exhaust. This layout allows for a single OCV  822  to simultaneously switch each of the four SRFF-1L rocker arm  1100  assemblies necessary for cylinder deactivation. Finally, the system is controlled electronically from the ECU  825  to the OCV  822  to switch between normal-lift mode and no-lift mode. 
     The engine layout for one exhaust and one intake valve using the SRFF-1L  1100  is shown in  FIG. 90 . The packaging of the SRFF-1L  1100  is similar to that of the standard valvetrain. The cylinder head requires modification to provide an oil feed from the lower gallery  805  to the OCV  822  ( FIGS. 88, 91 ). Additionally, a second (upper) oil gallery  802  is required to connect the OCV  822  and the switching ports  506  of the DFHLA  110 . The basic engine cylinder head architecture remains the same such that the valve centerline, camshaft centerline, and DFHLA  110  centerline remain constant. Because these three centerlines are maintained relative to a standard valvetrain, and because the SRFF-1L  1100  remains compact, the cylinder head height, length, and width remain nearly unchanged compared to a standard valvetrain system. 
     2. CDA-1L System Enabling Technologies 
     Several technologies used in this system have multiple uses in varied applications, they are described herein as components of the DVVL system disclosed herein. These include: 
     2.1. Oil Control Valve (OCV) 
     As described in earlier sections, and shown in  FIGS. 88, 91, 92, and 93 , an oil control valve (OCV)  822  is a control device that directs or does not direct pressurized hydraulic fluid to cause the rocker arm  1100  to switch between normal-lift mode and no-lift mode. The OCV is intelligently controlled, for example using a control signal sent by the ECU  825 . 
     2.2. Dual Feed Hydraulic Lash Adjustor (DFHLA) 
     Many hydraulic lash adjusting devices exist for maintaining lash in engines. For DVVL switching of rocker arm  100  ( FIG. 4 ), traditional lash management is required, but traditional HLA devices are insufficient to provide the necessary oil flow requirements for switching, withstand the associated side-loading applied by the assembly  100  during operation, and fit into restricted package spaces. A compact dual feed hydraulic lash adjuster  110  (DFHLA), used together with a switching rocker arm  100  is described, with a set of parameters and geometry designed to provide optimized oil flow pressure with low consumption, and a set of parameters and geometry designed to manage side loading. 
     As illustrated in  FIG. 10 , the ball plunger end  601  fits into the ball socket  502  that allows rotational freedom of movement in all directions. This permits side and possibly asymmetrical loading of the ball plunger end  601  in certain operating modes, for example when switching from high-lift to low-lift and vice versa. In contrast to typical ball end plungers for HLA devices, the DFHLA  110  ball end plunger  601  is constructed with thicker material to resist side loading, shown in  FIG. 11  as plunger thickness  510 . 
     Selected materials for the ball plunger end  601  may also have higher allowable kinetic stress loads, for example, chrome vanadium alloy. 
     Hydraulic flow pathways in the DFHLA  110  are designed for high flow and low pressure drop to ensure consistent hydraulic switching and reduced pumping losses. The DFHLA is installed in the engine in a cylindrical receiving socket sized to seal against exterior surface  511 , illustrated in  FIG. 11 . The cylindrical receiving socket combines with the first oil flow channel  504  to form a closed fluid pathway with a specified cross-sectional area. 
     As shown in  FIG. 11 , the preferred embodiment includes four oil flow ports  506  (only two shown) as they are arranged in an equally spaced fashion around the base of the first oil flow channel  504 . Additionally, two second oil flow channels  508  are arranged in an equally spaced fashion around ball end plunger  601 , and are in fluid communication with the first oil flow channel  504  through oil ports  506 . Oil flow ports  506  and the first oil flow channel  504  are sized with a specific area and spaced around the DFHLA  110  body to ensure even flow of oil and minimized pressure drop from the first flow channel  504  to the third oil flow channel  509 . The third oil flow channel  509  is sized for the combined oil flow from the multiple second oil flow channels  508 . 
     2.3. Sensing and Measurement 
     Information gathered using sensors may be used to verify switching modes, identify error conditions, or provide information analyzed and used for switching logic and timing. As can be seen, the sensing and measurement embodiments described in earlier sections pertaining to the DVVL system may also be applied to the CDA-1L system. Therefore, the valve position and/or motion sensing and logic used in DVVL, may also be used in the CDA system. Similarly, the sensing and logic used in determining the position/motion of the rocker arms, or the relative position/motion of the rocker arms relative to each other used for the DVVL system may also be used in the CDA system. 
     2.4. Torsion Spring Design and Implementation 
     A robust torsion spring  1124  design that provides more torque than conventional existing rocker arm designs, while maintaining high reliability, enables the CDA-1L system to maintain proper operation through all dynamic operating modes. The design and manufacture of the torsion springs  1124  are described in later sections. 
     3. Switching Control and Logic 
     3.1. Engine Implementation 
     CDA-1L embodiments may include any number of cylinders, for example 4 and 6 cylinder in-line and 6 and 8 cylinder V-configurations. 
     3.2. Hydraulic Fluid Delivery System to the Rocker Arm Assembly 
     As shown in  FIG. 91 , the hydraulic fluid system delivers engine oil at a controlled pressure to the CDA-1L switching rocker arm  1100 . In this arrangement, engine oil from the cylinder head  801  that is not pressure regulated feeds into the DFHLA  110  via the lower oil gallery  805 . This oil is always in fluid communication with the lower port  512  of the DFHLA  110 , where it is used to perform normal hydraulic lash adjustment. Engine oil from the cylinder head  801  that is not pressure regulated is also supplied to the oil control valve  822 . Hydraulic fluid from OCV  822 , supplied at a controlled pressure, is supplied to the upper oil gallery  802 . Switching of OCV  822  determines the lift mode for each of the CDA-1L rocker arm  1100  assemblies that comprise a CDA deactivation system  1400  for a given engine cylinder. As described in following sections, actuation of the OCV valve  822  is directed by the engine control unit  825  using logic based on both sensed and stored information for particular physical configuration, switching window, and set of operating conditions, for example, a certain number of cylinders and a certain oil temperature. Pressure regulated hydraulic fluid from the upper gallery  802  is directed to the DFHLA  110  upper port  506 , where it is transmitted to the switching rocker arm assembly  1100 . Hydraulic fluid is communicated through the rocker arm assembly  1100  to the latch pin  1202  assembly, where it is used to initiate switching between normal-lift and no-lift states. 
     Purging accumulated air in the upper gallery  802  is important to maintain hydraulic stiffness and minimize variation in the pressure rise time. Pressure rise time directly affects the latch movement time during switching operations. The passive air bleed port  832 , shown in  FIG. 91  was added to the high points in the upper gallery  802  to vent accumulated air into the cylinder head air space under the valve cover. 
     3.2.1. Hydraulic Fluid Delivery for Normal-Lift Mode 
       FIG. 92  shows the SRFF-1L  1100  in the default position where the electronic signal to the OCV  822  is absent, and also shows a cross section of the system and components that enable operation in normal-lift mode: OCV  822 , DFHLA  110 , latch spring  1204 , latch  1202 , outer arm  1102 , cam  1320 , roller bearing  1116 , inner arm  1108 , valve pad  1140  and engine valve  112 . Unregulated engine oil pressure in the lower gallery  805  is in communication with the lash compensation (lower) port  512  of the DFHLA  110  to enable standard lash compensation. The OCV  822  regulates oil pressure to the upper oil gallery  802 , which then supplies oil to the upper port  506  at 0.2 to 0.4 bar when the ECU  825  electrical signal is absent. This pressure value is below the pressure required to compress the latch spring  1204  move the latch pin  1202 . This pressure value serves to keep the oil circuit full of oil and free of air to achieve the required system response. The cam  1320  lobe contacts the roller bearing, rotating outer arm  1102  about the DFHLA  110  ball socket to open and close the valve. When the latch  1202  is engaged, the SRFF-1L functions similarly to a standard RFF rocker arm assembly. 
     3.2.2. Hydraulic Fluid Delivery for No-Lift Mode 
       FIGS. 93  A, B, and C show detailed views of the SRFF-1L  1100  during cylinder deactivation (no-lift mode). The Engine Control Unit (ECU)  825  ( FIG. 91 ) provides a signal to the OCV  822  such that oil pressure is supplied to the latch  1202  causing it to retract as shown in  FIG. 93 b   . The pressure required to fully retract the latch is 2 bar or greater. The higher torsion spring  1124  ( FIGS. 88, 99 ) preload in this single-lobe CDA embodiment enables the camshaft lobe  1320  to stay in contact with the inner arm  1108  roller bearing  1116  as this occurs in lost motion, and the engine valve remains closed as shown in  FIG. 93   c.    
     3.3. Operating Parameters 
     An important factor in operating a CDA system  1400  ( FIG. 91 ) is the reliable control of switching between normal-lift mode to no-lift mode. CDA valve actuation systems  1400  can only be switched between modes during a predetermined window of time. As described above, switching from high-lift mode to low-lift mode and vice versa is initiated by a signal from the engine control unit (ECU)  825  ( FIG. 91 ) using logic that analyzes stored information, for example a switching window for particular physical configuration, stored operating conditions, and processed data that is gathered by sensors. Switching window durations are determined by the CDA system physical configuration, including the number of cylinders, the number of cylinders controlled by a single OCV, the valve lift duration, engine speed, and the latch response times inherent in the hydraulic control and mechanical system. 
     3.3.1. Gathered Data 
     Real-time sensor information includes input from any number of sensors, as illustrated in the exemplary CDA-1L system  1400  illustrated in  FIG. 91 . As described previously, sensors may include 1) valve stem movement  829 , as measured in one embodiment using a linear variable differential transformer (LVDT), 2) motion/position  828  and latch position  827  using a Hall-effect sensor or motion detector, 3) DFHLA movement  826  using a proximity switch, Hall effect sensor, or other means, 4) oil pressure  830 , and 5) oil temperature  890 . Cam shaft rotary position and speed may be gathered directly or inferred from the engine speed sensor. 
     In a hydraulically actuated VVA system, the oil temperature affects the stiffness of the hydraulic system used for switching in systems such as CDA and VVL. If the oil is too cold, its viscosity slows switching time, causing a malfunction. This temperature relationship is illustrated for an exemplary CDA-1L switching rocker arm  1100  system  1400  in  FIG. 96 . An accurate oil temperature, in one embodiment taken with a sensor  890  shown in  FIG. 91 , located near the point of use rather than in the engine oil crankcase, provides accurate information. In one example, the oil temperature in a CDA system  1400 , monitored close to the oil control valves (OCV)  822 , must be greater than or equal to 20 degrees C. to initiate no-lift (unlatched) operation with the required hydraulic stiffness. Measurements can be taken with any number of commercially available components, for example a thermocouple. The oil control valves are described further in published US Patent Applications US2010/0089347 published Apr. 15, 2010 and US2010/0018482 published Jan. 28, 2010 both hereby incorporated by reference in their entirety. 
     Sensor information is sent to the Engine Control Unit (ECU)  825  as a real-time operating parameter. 
     3.4. Stored Information 
     3.4.1. Switching Window Algorithms 
     The SRFF requires mode switching from the normal-lift to no-lift (deactivated), state and vice-versa. Switching is required to occur in less than one camshaft revolution to ensure proper engine operation. Mode switching can occur only when the SRFF is on the base circle  1322  ( FIG. 101 ) of the cam  1320 . Switching between valve lift states cannot occur when the latch  1202  ( FIG. 93 ) is loaded and movement is restricted. The latch  1202  transition period between full and partial engagement must be controlled to keep the latch  1202  from slipping. Switching windows combined with electro-mechanical latch response times inherent in the CDA system  1400  ( FIG. 91 ) identify the opportunities for mode switching. 
     The intended functional parameters of the SRFF based CDA system  1400  is analogous to the Type-V switching roller lifter designs that are in production today. The mode switch between normal-lift and no-lift is set to occur during the base circle  1322  event and be synchronized to the camshaft  1300  rotational position. The SRFF default position is set to normal-lift. The oil flow demand on the SRFF is also similar to the Type-V CDA production systems. 
     A critical shift is defined as an unintended event that may occur when latch is partially engaged, causing the valve to lift partially and suddenly drop back to the valve seat. This condition is unlikely, when the switching commands are executed during prescribed parameters of oil temperature, engine speeds with the camshaft position synchronized switching. The critical shift event creates an impact load to the DFHLA  110 , which may require high strength DFHLA&#39;s, described in earlier sections, as enabling system components. 
     The fundamentals the synchronized switching for the CDA system  1400  are illustrated in  FIG. 94 . The exhaust valve profile  1450  and intake valve profile  1452  are plotted as a function of crankshaft angle. The required switching window is defined as the sum of the time it takes for the following operations: 1) the OCV  822  valve to supply pressurized oil, 2) the hydraulic system pressure to overcome the biasing spring  1204  and cause latch  1202  mechanical movement, and 3) the complete movement of latch  1202  necessary for mode change from no-lift to normal-lift and visa-versa. Switching window duration  1454 , in this exhaust example, exists once the exhaust closes until the exhaust starts to open again. The latch  1202  remains restricted during the exhaust lift event. The timing windows that may cause critical shift  1456 , described in more detail in later sections, are identified in  FIG. 94 . The switching window for the intake can be described in similar terms relative to the intake lift profile. 
     Latch Pre-Load 
     The CDA-1L rocker arm  1100  switching mechanism is designed such that hydraulic pressure can be applied to the latch  1202  after the latch lash is absorbed, resulting in no change in function. This design parameter allows hydraulic pressure to be initiated by the OCV  822  in the upper oil gallery  802  during the intake valve lift event. Once the intake valve lift profile  1452  returns to the base circle  1322  no-load condition, the latch completes its movement to the specified latched or unlatched mode. This design parameter helps to maximize the available switching window. 
     Hydraulic Response Time Versus Temperature 
       FIG. 96  shows the dependence of latch  1202  response time on oil temperature using SAE 5W-30 oil. The latch  1202  response time, reflects the duration for the latch  1202  to move from normal-lift (latched) to no-lift (unlatched) position, and vice-versa. The latch  1202  response time requires ten milliseconds with an oil temperature of 20° C. and 3 bar oil pressure in the switching pressure port  506 . Latch response time is reduced to five milliseconds under the same pressure conditions at higher operating temperatures, for example 40° C. Hydraulic response times are used to determine switching windows. 
     Variable Valve Timing 
     Now, with reference to  FIGS. 94 and 95 , some camshaft drive systems are designed to have greater phasing authority/range of motion, relative to the crankshaft angle than standard drive systems. This technology may be referred to as variable valve timing, and must be considered along with engine speed when determining the allowable switching window duration  1454 . 
     The plots of valve lift profile as a function of crankshaft angle are shown in  FIG. 95 , illustrating the effect that variable valve timing has on the switching window duration  1454 . Exhaust valve lift profile  1450  and intake valve lift profile  1452  show a typical cycle with no variable valve timing capability that results in no switching window  1455  (also seen in  FIG. 94 ), Exhaust valve lift profile  1460  and intake valve lift profile  1462  show a typical cycle that has variable valve timing capability that results in no switching window  1464 . This example of variable valve timing results in an increase in the duration of the no switching window  1458 . Assuming a variable valve timing capability of 120 degrees crankshaft angle duration between the exhaust and intake camshafts, the time duration shift  1458  is 6 milliseconds at 3500 engine rpm. 
       FIG. 97  is a plot showing calculated and measured variations in switching time due to the effects of temperature and cam phasing. The plot is based on a switching window that ranges from 420 crankshaft degrees with camshaft phasing at minimum overlap  1468  to 540 crankshaft degrees with camshaft phasing at maximum overlap  1466 . The latch response time of 5 milliseconds shown on this plot is for normal engine operating temperatures of 40-120° C. The hydraulic response variation  1470  is measured from ECU  825  switching signal initiation until the hydraulic pressure is sufficient to cause the latch  1202  to move. Based on CDA system  1400  studies that use OCVs to control hydraulic oil pressure, the maximum variation is approximately 10 milliseconds. This hydraulic response variation  1470  takes into consideration voltage to the OCV  822 , temperature, and oil pressure in the engine. The phasing position with minimum overlap  1468  provides an available switching time of 20 milliseconds at 3500 engine rpm, and the total latch response time is 15 milliseconds, representing a 5 millisecond margin between the time available for switching and the latch  1202  response time. 
       FIG. 98  is also a plot showing calculated and measured variations in switching time due to the effects of temperature and cam phasing. The plot is based on a switching window that ranges from 420 crankshaft degrees with camshaft phasing at minimum overlap  1468  to 540 crankshaft degrees with camshaft phasing at maximum overlap  1466 . The latch response time of 10 milliseconds shown on this plot is for a cold engine operating temperatures of 20° C. The hydraulic response variation  1470  is measured from ECU  825  switching signal initiation until the hydraulic pressure is sufficient to cause the latch  1202  to move. Based on CDA system  1400  studies that use OCVs to control hydraulic oil pressure, the maximum variation is approximately 10 milliseconds. This hydraulic response variation  1470  takes into consideration voltage to the OCV  822 , temperature, and oil pressure in the engine. The phasing position with minimum overlap  1468  provides an available switching time of 20 milliseconds at 3500 engine rpm, and the total latch response time is 20 milliseconds, representing reduced design margin between the time available for switching and the latch  1202  response time. 
     3.4.2. Stored Operating Parameters 
     These variables include engine configuration parameters such as variable valve timing and predicted latch response times as a function of operating temperature. 
     3.5. Control Logic 
     As noted above, CDA switching can only occur during a small predetermined window of time under certain operating conditions, and switching the CDA system outside of the timing window may result in a critical shift event, that could result in damage to the valve train and/or other engine parts. Because engine conditions such as oil pressure, temperature, emissions, and load may vary rapidly, a high-speed processor can be used to analyze real-time conditions, compare them to known operating parameters that characterize a working system, reconcile the results to determine when to switch, and send a switching signal. These operations can be performed hundreds or thousands of times per second. In embodiments, this computing function may be performed by a dedicated processor, or by an existing multi-purpose automotive control system referred to as the engine control unit (ECU). A typical ECU has an input section for analog and digital data, a processing section that includes a microprocessor, programmable memory, and random access memory, and an output section that might include relays, switches, and warning light actuation. 
     In one embodiment, the engine control unit (ECU)  825  shown in  FIG. 91  accepts input from multiple sensors such as valve stem movement  829 , motion/position  828 , latch position  827 , DFHLA movement  826 , oil pressure  830 , and oil temperature  890 . Data such as allowable operating temperature and pressure for given engine speeds and switching windows are stored in memory. Real-time gathered information is then compared with stored information and analyzed to provide the logic for ECU  825  switching timing and control. 
     After input is analyzed, a control signal is transmitted by the ECU  825  to the OCV  822  to initiate switching operation, which may be timed to avoid critical shift events while meeting engine performance goals such as improved fuel economy and lowered emissions. If necessary, the ECU  825  may also alert operators to error conditions. 
     4. CDA-1L Rocker Arm Assembly 
       FIG. 99  illustrates a perspective view of an exemplary CDA-1L rocker arm  1100 . The CDA-1L rocker arm  1100  is shown by way of example only and it will be appreciated that the configuration of the CDA-1L rocker arm  1100  that is the subject of this application is not limited to the configuration of the CDA-1L rocker arm  1100  illustrated in the figures contained herein. 
     As shown in  FIGS. 99 and 100 , the CDA-1L rocker arm  1100  includes an outer arm  1102  having a first outer side arm  1104  and a second outer side arm  1106 . An inner arm  1108  is disposed between the first outer side arm  1104  and second outer side arm  1106 . The inner arm  1108  has a first inner side arm  1110  and a second inner side arm  1112 . The inner arm  1108  and outer arm  1102  are both mounted to a pivot axle  1114 , located adjacent the first end  1101  of the rocker arm  1100 , which secures the inner arm  1108  to the outer arm  1102  while also allowing a rotational degree of freedom pivoting about the pivot axle  1114  when the rocker arm  1100  is in a no-lift state. In addition to the illustrated embodiment having a separate pivot axle  1114  mounted to the outer arm  1102  and inner arm  1108 , the pivot axle  1114  may be integral to the outer arm  1102  or the inner arm  1108 . 
     The CDA-1L rocker arm  1100  has a bearing  1190  comprising a roller  1116  that is mounted between the first inner side arm  1110  and second inner side arm  1112  on a bearing axle  1118  that, during normal operation of the rocker arm, serves to transfer energy from a rotating cam (not shown) to the rocker arm  1100 . Mounting the roller  1116  on the bearing axle  1118  allows the bearing  1190  to rotate about the axle  1118 , which serves to reduce the friction generated by the contact of the rotating cam with the roller  1116 . As discussed herein, the roller  1116  is rotatably secured to the inner arm  1108 , which in turn may rotate relative to the outer arm  1102  about the pivot axle  1114  under certain conditions. In the illustrated embodiment, the bearing axle  1118  is mounted to the inner arm  1108  in the bearing axle apertures  1260  of the inner arm  1108  and extends through the bearing axle slots  1126  of the outer arm  1102 . Other configurations are possible when utilizing a bearing axle  1118 , such as having the bearing axle  1118  not extend through bearing axle slots  1126  but still mounted in bearing axle apertures  1260  of the inner arm  1108 , for example. 
     When the rocker arm  1100  is in a no-lift state, the inner arm  1108  pivots downwardly relative to the outer arm  1102  when the lifting portion of the cam ( 1324  in  FIG. 101 ) comes into contact with the roller  1116  of bearing  1190 , thereby pressing it downward. The axle slots  1126  allow for the downward movement of the bearing axle  1118 , and therefore of the inner arm  1108  and bearing  1190 . As the cam continues to rotate, the lifting portion of the cam rotates away from the roller  1116  of bearing  1190 , allowing the bearing  1190  to move upwardly as the bearing axle  1118  is biased upwardly by the bearing axle torsion springs  1124 . The illustrated bearing axle springs  1124  are torsion springs secured to mounts  1150  located on the outer arm  1102  by spring retainers  1130 . The torsion springs  1124  are secured adjacent the second end  1103  of the rocker arm  1100  and have spring arms  1127  that come into contact with the bearing axle  1118 . As the bearing axle  1118  and spring arm  1127  move downward, the bearing axle  1118  slides along the spring arm  1127 . The configuration of rocker arm  1100  having the torsion springs  1124  secured adjacent the second end  1103  of the rocker arm  1100 , and the pivot axle  1114  located adjacent the first end  1101  of the rocker arm, with the bearing axle  1118  between the pivot axle  1114  and the axle spring  1124 , lessens the mass near the first end  1101  of the rocker arm. 
     As shown in  FIGS. 101 and 102 , the valve stem  1350  is also in contact with the rocker arm  1100  near its first end  1101 , and thus the reduced mass at the first end  1101  of the rocker arm  1100  reduces the mass of the overall valve train (not shown), thereby reducing the force necessary to change the velocity of the valve train. It should be noted that other spring configurations may be used to bias the bearing axle  1118 , such as a single continuous spring. 
       FIG. 100  illustrates an exploded view of the CDA-1L rocker arm  1100  of  FIG. 99 . The exploded view in  FIG. 100  and the assembly view in  FIG. 99 , show bearing  1190 , a needle roller-type bearing that comprises a substantially cylindrical roller  1116  in combination with needles  1200 , which can be mounted on a bearing axle  1118 . The bearing  1190  serves to transfer the rotational motion of the cam to the rocker arm  100  that in turn transfers motion to the valve stem  1350 , for example in the configuration shown in  FIGS. 101 and 102 . As shown in  FIGS. 99 and 100 , the bearing axle  1118  may be mounted in the bearing axle apertures  1260  of the inner arm  1108 . In such a configuration, the axle slots  1126  of the outer arm  1102  accept the bearing axle  1118  and allow for lost motion movement of the bearing axle  1118  and by extension the inner arm  1108  when the rocker arm  1100  is in a non-lift state. “Lost motion” movement can be considered movement of the rocker arm  1100  that does not transmit the rotating motion of the cam to the valve. In the illustrated embodiments, lost motion is exhibited by the pivotal motion of the inner arm  1108  relative to the outer arm  1102  about the pivot axle  1114 . 
     Other configurations other than bearing  1190  also permit the transfer of motion from the cam to the rocker arm  1100 . For example, a smooth non-rotating surface (not shown) for interfacing with the cam lift lobe ( 1320  in  FIG. 101 ) may be mounted on or formed integral to the inner arm  1108  at approximately the location where the bearing  1190  is shown in  FIG. 99  relative to the inner arm  1108  and rocker arm  1100 . Such a non-rotating surface may comprise a friction pad formed on the non-rotating surface. In another example, alternative bearings, such as bearings with multiple concentric rollers, may be used effectively as a substitute for bearing  1190 . 
     With reference to  FIGS. 99 and 100 , the elephant foot  1140  is mounted on the pivot axle  1114  between the first  1110  and second  1112  inner side arms. The pivot axle  1114  is mounted in the inner pivot axle apertures  1220  and outer pivot axle apertures  1230  adjacent the first end  1101  of the rocker arm  1100 . Lips  1240  formed on inner arm  1108  prevent the elephant foot  1140  from rotating about the pivot axle  1114 . The elephant foot  1140  engages the end of the valve stem  1350  as shown in  FIG. 102 . In an alternative embodiment, the elephant foot  1140  may be removed, and instead an interfacing surface complementary to the tip of the valve stem  1350  may be placed on the pivot axle  1114 . 
       FIGS. 101 and 102  illustrate a side view and front view, respectively, of rocker arm  1100  in relation to a cam  1300  having a lift lobe  1320  with a base circle  1322  and lifting portion  1324 . A roller  1116  is illustrated in contact with the lift lobe  1320 . A dual feed hydraulic lash adjuster (DFHLA)  110  engages the rocker arm  1100  adjacent its second end  1103 , and applies upward pressure to the rocker arm  1100 , and in particular the outer rocker arm  1102 , while mitigating against valve lash. The valve stem  1350  engages the elephant foot  1140  adjacent the first end  1101  of the rocker arm  1100 . In the normal-lift state, the rocker arm  1100  periodically pushes the valve stem  1350  downward, which serves to open the corresponding valve (not shown). 
     4.1. Torsion Spring 
     As described in following sections, a rocker arm  1100  in the no-lift state may be subjected to excessive pump-up of the lash adjuster  110 , whether due to excessive oil pressure, the onset of non-steady-state conditions, or other causes. This may result in an increase in the effective length of the lash adjuster  110  as pressurized oil fills its interior. Such a scenario may occur for example during a cold start of the engine, and could take significant time to resolve on its own if left unchecked and could even result in permanent engine damage. Under such circumstances, the latch  1202  may not be able to activate the rocker arm  1100  until the lash adjuster  110  has returned to a normal operating length. In this scenario, the lash adjuster  110  applies upward pressure to the outer arm  1102 , bringing the outer arm  1102  closer to the cam  1300 . 
     The lost motion torsion spring  1124  on the SRFF-1L was designed to provide sufficient force to keep the roller bearing  1116  in contact with the camshaft lift lobe  1320  during no-lift operation to ensure controlled acceleration and deceleration of the inner arm subassembly and controlled return of the inner arm  1108  to the latching position while preserving the latch lash. A pump-up scenario requires a stronger torsion spring  1124  to compensate for the additional force from pump-up. 
     Rectangular wire cross sections for the torsion springs  1124  were used to reduce the package space, keeping the assembly moment of inertia low and providing sufficient cross section height to sustain the operating loads. Stress calculations and FEA, and test validation, described in following sections, were used in developing the torsion spring  1124  components. 
     A torsion spring  1124  ( FIG. 99 ) design and manufacturing process is described that results in a compact design with a generally rectangular shaped wire made with selected materials of construction. 
     Now, with reference to  FIGS. 30A, 30B, and 99 , the torsion spring  1124  is constructed from a wire  397  that is generally trapezoidal in shape. The trapezoidal shape is designed to allow wire  397  to deform into a generally rectangular shape as force is applied during the winding process. After torsion spring  1124  is wound, the shape of the resulting wires can be described as similar to a first wire  396  with a generally rectangular shape cross section.  FIG. 99  shows two torsion spring embodiments, illustrated as multiple coils  398 ,  399  in cross section. In a preferred embodiment, wire  396  has a rectangular cross sectional shape, with two elongated sides, shown here as the vertical sides  402 ,  404  and a top  401  and bottom  403 . The ratio of the average length of side  402  and side  404  to the average length of top  401  and bottom  403  of the coil can be any value less than 1. This ratio produces more stiffness along the coil axis  419  of bending than a spring coiled with round wire with a diameter equal to the average length of top  401  and bottom  403  of the coil  398 . In an alternate embodiment, the cross section wire shape has a generally trapezoidal shape with a larger top  401  and a smaller bottom  403 . 
     In this configuration, as the coils are wound, elongated side  402  of each coil rests against the elongated side  402  of the previous coil, thereby stabilizing the torsion springs  1124 . The shape and arrangement holds all of the coils in an upright position, preventing them from passing over each other or angling when under pressure. 
     When the rocker arm assembly  1100  is operating, the generally rectangular or trapezoidal shape of the torsion springs  1124 , as they bend about axis  419  shown in  FIGS. 30A and 30B , produce high part stress, particularly tensile stress on top surface  401 . To meet durability requirements, a combination of techniques and materials are used together. For example, the torsion spring may be made of a material that includes Chrome Vanadium alloy steel along with this design to improve strength and durability. The torsion spring may be heated and quickly cooled to temper the springs. This reduces residual part stress. Impacting the surface of the wire  396 ,  397  used for creating the torsion springs with projectiles, or ‘shot peening’ is used to put residual compressive stress in the surface of the wire  396 ,  397 . The wire  396 ,  397  is then wound into the torsion spring. Due to their shot peening, the resulting torsion springs can now accept more tensile stress than identical springs made without shot peening. 
     4.2. Torsion Spring Pocket 
     As illustrated in  FIG. 100 , knob  1262  extends from the end of the bearing axle  1118  and creates a slot  1264  in which the spring arm  1127  sits. In one alternative, a hollow bearing axle  1118  may be used along with a separate spring mounting pin (not shown) comprising a feature such as the knob  1262  and slot  1264  for mounting the spring arm  1127 . 
     4.3. Outer Arm Assembly 
     4.3.1. Latch Mechanism Description 
     The mechanism for selectively deactivating the rocker arm  1100 , which in the illustrated embodiment is found near the second end  1103  of the rocker arm  1100 , is shown in FIG.  100  as comprising latch  1202 , latch spring  1204 , spring retainer  1206  and clip  1208 . The latch  1202  is configured to be mounted inside the outer arm  1102 . The latch spring  1204  is placed inside the latch  1202  and secured in place by the latch spring retainer  1206  and clip  1208 . Once installed, the latch spring  1204  biases the latch  1202  toward the first end  1101  of the rocker arm  1100 , allowing the latch  1202 , and in particular the engaging portion  1210  to engage the inner arm  1108 , thereby preventing the inner arm  1108  from moving with respect to the outer arm  1102 . When the latch  1202  is engaged with the inner arm in this way, the rocker arm  1100  is in the normal-lift state, and will transfer motion from the cam to the valve stem. 
     In the assembled rocker arm  1100 , the latch  1202  alternates between normal-lift and no-lift states. The rocker arm  1100  may enter the no-lift state when oil pressure sufficient to counteract the biasing force of latch spring  1204  is applied, for example, through the port  1212  which is configured to permit oil pressure to be applied to the surface of the latch  1202 . When the oil pressure is applied, the latch  1202  is pushed toward the second end  1103  of the rocker arm  1100 , thereby withdrawing the latch  1202  from engagement with the inner arm  1108  and allowing the inner arm  1108  to rotate about the pivot axle  1114 . In both the normal-lift and no-lift states, the linear portion  1250  of orientation clip  1214  engages the latch  1202  at the flat surface  1218 . The orientation clip  1214  is mounted in the clip apertures  1216 , and thereby maintains a horizontal orientation of the linear portion  1250  relative to the rocker arm  1100 . This restricts the orientation of the flat surface  1218  to also be horizontal, thereby orienting the latch  1202  in the appropriate direction for consistent engagement with the inner arm  1108 . 
     4.3.2. Latch Pin Design 
     As shown in  FIGS. 93  A, B, C, the SRFF-1L rocker arm  1100  latch  1202  operating in no-lift mode is retracted inside the outer arm  1202 , while the inner arm  1108  follows the camshaft lift lobe  1320 . Under certain conditions, transitioning from no-lift mode to normal-lift mode can result in a condition shown in  FIG. 103 , where the latch  1202  extends before the inner arm  1108  returns to the position where the latch  1202  normally engages. 
     A re-engagement feature was added to the SRFF to prevent the condition where the inner arm  1108  is blocked and trapped in a position below the latch  1202 . An inner arm sloped surface  1474  and a latch sloped surface  1472  were optimized to provide smooth latch  1202  movement to the retracted position when the inner arm  1108  contacts the latch sloped surface  1472 . The design avoids damage to latch mechanism that may be caused by pressure changes at the switching pressure port  506  ( FIG. 88 ). 
     4.4. System Packaging 
     The SRFF-1F design is focused on minimizing valvetrain packaging changes compared to a standard production layout. Important design parameters include relative placement of the camshaft lobes in relation to the SRFF roller bearing, and axial alignment between the steel camshaft and aluminum cylinder head. The steel and aluminum components have different thermal growth coefficients that can shift the camshaft lobes relative to the SRFF-1F. 
       FIG. 104  shows both proper and poor alignment of the single camshaft lobe relative to the SRFF-1L  1100  outer arm  1102  and bearing  1116 . The proper alignment shows the camshaft lift lobe  1320  centered over the roller bearing  1116 . The single camshaft lobe  1320  and SRFF-1L  1110  is designed to avoid edge loading  1482  on the roller bearing  1116  and avoid cam lobe  1320  contact  1480  with the outer arm  1102 . The elimination of camshaft no-lift lobes found in multi-lobe CDA configurations relaxes the requirements for tight manufacturing tolerances and assembly control of the camshaft lobe width and position, making the camshaft manufacturing process similar to that of standard camshafts used on Type II engines. 
     4.5. CDA-1L Latch Mechanism Hydraulic Operation 
     As previously mentioned, pump-up is a term used to describe a condition in which the HLA is extended past its intended working dimension; thereby preventing the valve from returning to its seat during the base circle event. 
       FIG. 105  below shows a standard valvetrain system and the forces acting on the roller finger follower assembly (RFF)  1496  during a camshaft base circle event. The hydraulic lash adjuster force  1494  is a combination of the hydraulic lash adjuster (HLA)  1493  force generated by the oil pressure in the lash compensation port  1491  and the HLA internal spring force. The cam reaction force  1490  is between the camshaft  1320  and the RFF bearing. The reaction force  1492  is between the RFF  1496  and the valve  112  tip. The force balance must be such that the valve spring force  1492  will prevent unintentional opening of the valve  112 . If the valve reaction force  1492  generated by the HLA force  1494  and cam reaction force  1490  exceeds the seating force required to seat the valve  112 , then the valve  112  will be lifted and held open during base circle operation, which is undesirable. This description of the standard fixed arm system does not include the dynamic operating loads. 
     The SRFF-1L  1100  was designed with additional consideration for pump-up when the system is in no-lift mode. Pump-up of the DFHLA  110  when the SRFF-1L  1100  is in no-lift mode can create a condition in which the inner arm  1108  does not return to the position where the latch  1202  can re-engage the inner arm  1108 . 
     The SRFF-1L  1100  reacts similarly to a standard RFF  1496  ( FIG. 105 ) when the SRFF-1L  1100  is in normal-lift mode. Maintaining the required latch lash to switch the SRFF-1L  1100  while preventing pump-up is resolved by applying additional force from the torsion springs  1124  to overcome the HLA force  1494  in addition to the torsional already force required to return the inner arm  1108  to its the latch engagement position. 
       FIG. 106  shows the balance of forces acting on the SRFF-1L  1100  when the system is in no-lift mode: the DFHLA force  1499 , caused by the oil pressure at the lash compensator port  512  ( FIG. 88 ) plus the plunger spring force  1498 , the cam reaction force  1490 , and the torsion spring force  1495 . The torsion force  1495  produced by springs  1124  is converted, via the bearing axle  1118  and the spring arms  1127 , to spring reaction force  1500  acting on the inner arm  1108 . 
     The torsion springs  1124  in the SRFF-1L rocker arm assembly  1100  were designed to provide sufficient force to keep the roller bearing  1116  in contact with the camshaft lift lobe  1320  during no-lift mode to ensure controlled acceleration and deceleration of the inner arm  1108  subassembly and return the inner arm  1108  to the latching position while preserving the latch lash  1205 . The torsion spring  1124  design for SRFF-1L  1100  design also accounts for a variation in oil pressure at the lash compensation port  512  when the system is in no-lift mode. Oil pressure regulation can reduce the load requirements for the torsion springs  1124  with direct effect on the spring sizing. 
       FIG. 107  shows the requirements for oil pressure in the lash compensation pressure port  512 . Limited oil pressure for the SRFF-1L is only required when the system is in no-lift mode. Consideration for synchronized switching, described in earlier sections, limits the no-lift mode for temperatures lower than 20° C. 
     4.6. CDA-1L Assembly Lash Management 
       FIG. 108  shows the latch lash  1205  for the SRFF-1L  1100 . For a single-lobe CDA system, the total mechanical lash  1505  is reduced to a single latch lash  1205  value, as opposed to the sum of camshaft lash  1504  and latch lash  1205  for CDA designs with more than one lobe. The latch lash  1205  for the SRFF-1L  1100  is the distance between the latch  1202  and the inner arm  1108 . 
       FIG. 109  compares the opening ramp on a camshaft designed for a three-lobe SRFF and the single-lobe SRFF-1L. 
     Camshaft lash was eliminated by design for the single-lobe SRFF-1L. The elimination of the camshaft lash  1504  allows further optimization of the camshaft lift profile, by creating a lifting ramp reduction  1510 , thus allowing for longer lift events. The camshaft opening ramps  1506  for the SRFF-1L are reduced up to 36% from the camshaft opening ramps  1506  required for similar designs using multiple lobes. 
     In addition, mechanical lash variation on the SRFF-1L is improved 39% over an analogous three-lobe design due to the elimination of the camshaft lash and the features associated with it, for example, manufacturing tolerances for the camshaft no-lift lobes base circle radius, lobe run-out, required slider pad to slider pad and slider pad to roller bearing parallelism. 
     4.7. CDA-1L Assembly Dynamics 
     4.7.1. Detailed Description 
     The SRFF-1L rocker arm  1100  and system  1400  ( FIG. 91 ) is designed to meet the dynamic stability requirements for the entire engine operating range. SRFF stiffness and moment of inertia (MOI) were analyzed for the SRFF design. The MOI of the SRFF-1L assembly  1100  is measured about the pivot axle  1114  ( FIG. 99 ) which is the rotational axis that passes through the SRFF socket that is in contact with the DFHLA  110 . Stiffness is measured at the interface between cam  1320  and bearing  1116 .  FIG. 110  shows measured stiffness plotted against calculated assembly MOI. The SRFF-1L relationship between stiffness and MOI compares well with standard RFF&#39;s used on Type II engines currently in production. 
     4.7.2. Analysis 
     Several design and Finite Element Analysis (FEA) iterations were performed to maximize the stiffness and reduce MOI over the DFHLA end of the SRFF. The mass intensive components were placed over the DFHLA end of the SRFF to minimize the MOI. The torsion springs  1124 , one of the heaviest components in the SRFF assembly were positioned in close proximity to the SRFF rotational axis. The latching mechanism was also located near the DFHLA. The vertical section height of the SRFF was increased to maximize stiffness while minimizing MOI. 
     The SRFF designs were optimized using load information from kinematic modeling. Key input parameters for the analysis include valvetrain layout, SRFF elements of mass, moment of inertia, stiffness (predicted by the FEA), mechanical lash, valve spring loads and rates, DFHLA geometry and plunger spring, and valve lift profiles. Next, the system was altered to meet the predicted dynamic targets, by optimizing the stiffness versus the effective mass over the valve of the CDA SRFF. The effective mass over the valve represents the ratio between the MOI in respect to the pivot point of the SRFF and the square distance between the valve and the SRFF pivot. The tested dynamic performance is described in later sections. 
     5. Design Verification and Testing 
     5.1. Valve Train Dynamic Results 
     Dynamic behavior of a valvetrain is important in controlling the Noise Vibration and Harshness (NVH) while meeting the durability and performance targets of an engine. Valvetrain dynamics are partially influenced by the stiffness and MOI of the SRFF component. The MOI of the SRFF can be readily calculated and the stiffness is estimated through Computer Aided Engineering (CAE) techniques. Dynamic valve motion is also influenced by a variety of factors, so tests were conducted gain assurance in high speed valve control. 
     A motorized engine test rig was utilized for valvetrain dynamics. A cylinder head was instrumented prior to the test. Oil was heated to represent actual engine conditions. A speed sweep was performed from idle speed to 7500 rpm, recording data as defined by engine speed. Dynamic performance was determined by evaluating valve closing velocity and valve bounce. The SRFF-1L was strain gaged for the purpose of monitoring load. Valve spring loads were held constant to the fixed system for consistency. 
       FIG. 111  illustrates the resultant seating closing velocity of an intake valve. Data was acquired for eight consecutive events showing the minimum, average, and maximum velocities relative to engine speed. The target velocity is shown as the maximum speed for seating velocity that is typical in the industry. The target seating velocity was maintained up to approximately 7500 engine rpm which illustrates acceptable dynamic control for passenger car engine applications. 
     5.2. Torsion Spring Validation 
     Torsion springs are key components for the SRFF-1L design, especially during high speed operation. Concept validation was conducted on the springs to validate the robustness. Three elements of the spring design were tested for proof of concept. First, load loss was documented under the conditions of high cycling at operating temperature. Spring load loss, or relaxation, represents the reduction of the spring load at end of test from beginning of test. The load loss was also documented by applying highest stress levels and subjecting parts to high temperatures. Second, the durability and the springs were tested at worst case load and cycled to validate fatigue life, as well as the load loss as mentioned. Finally, the function of the lost motion springs were validated by using lowest load springs and verifying that the DFHLA does not pump up during all operating conditions in CDA mode. 
     The torsion springs were cycled at engine operating temperatures in the engine oil environment on a targeted fixture test. Torsion springs were cycled with the full stroke of the application with the highest preload conditions to represent worst case stress. The cycling target value was set at 25 million and 50 million cycles. Torsion springs were also subjected to a heat-set test in which they were loaded to highest application stress and held at 140° C. for 50 hours and measured for load loss. 
       FIG. 112  summarizes the load loss for both the cycling test and the heat set test. All parts passed with a maximum load loss of 8% while the design target was set to 10% maximum load loss. 
     The results indicated a maximum load loss of 8% and met the design target. Many of the tests showed minimal load loss near 1%. All tests were safely within the design guidelines for load loss. 
     5.3. Pump-Up Robustness During Cylinder Deactivation 
     Torsion springs  1124  ( FIG. 99 ) are designed to prevent the HLA pump-up to preserve the latch lash  1205  ( FIG. 108 ) when the system operates in no-lift mode. The test apparatus was designed to sustain engine oil pressure at the lash compensation pressure port over the range of oil temperatures and engine speed conditions where mode switching is required. 
     Validation experiments were performed to prove torsion spring  1124  ability to preserve latch lash  1205  at required conditions. The tests were conducted on motorized engines, with instrumentation for measuring the valve and the CDA SRFF motion, oil pressure and temperature at the lash compensation pressure port  512  ( FIG. 88 ) and switching pressure port  506  ( FIG. 88 ). 
     Low limit lost motion springs were used to simulate worst condition. This test was conducted at 3500 rpm which represents the maximum switching speed. Two operating temperatures were considered of 58° C. and 130° C. Test results show pump-up at pressures 25% higher than the application requirement. 
       FIG. 113  shows the lowest pump-up pressure measured  1540 , which is on the exhaust side at 58° C. Pump-up pressure for the intake at 58° C. and 130° C. and exhaust at 130° C. were higher than the pump-up pressure of the exhaust side at 58° C. The SRFF was in switching mode, having events on normal-lift and events in no-lift mode. Proximity probes were used to detect valve motion in order to validate the SRFF mode state at corresponding pressure at the switching pressure port  506 . The pressure in the lash compensator port  512  was gradually increased and switching from no-lift mode to normal-lift mode was monitored. The pressure at which the system ceased to switch was recorded as pump-up pressure  1540 . The system safely avoids pump-up pressures when the oil pressure is maintained at or below 5 bar for the SRFF-1L design. Concept testing was conducted with specially procured high limit torque torsion spring to simulate the worst case fatigue design margin condition. The concept testing conducted on the high load torsion spring met the required design goal. 
     5.4. Validation of Mechanical Lash During Switching Durability 
     Mechanical lash control is important to valvetrain dynamic stability and must be maintained through the life of the engine. A test with loading of the latch and switching between normal-lift mode and no-lift mode was considered appropriate to validate the wear and the performance of the latch mechanism. Switching durability was tested by switching the latch from the engaged to disengaged position, cycling the SRFF in no-lift mode, engaging the latch with the inner arm and cycling the SRFF in normal-lift mode. One cycle is defined to disengage and then re-engage the latch and exercise the SRFF in the two modes. The durability target for switching is 3,000,000 cycles. 3,000,000 cycles represents the equivalent of one engine life. One engine life is defined as an equivalent of 200,000 miles which is safely above the 150,000 mile standard. Parts were tested at highest switching speed target of 3500 engine rpm to simulate worst case dynamic load during switching. 
       FIG. 114  illustrates the change in mechanical lash at periodic inspection points during the test. This test was conducted on one bank of a six cylinder engine fixture. Since there are three cylinders per bank and four SRFF-1L&#39;s per cylinder, twelve profiles are shown. The mechanical lash limit change of 0.020 mm was established as the design wear target. All SRFF-1L&#39;s show a safe margin of lash wear below the wear target at the equivalent of the vehicle life. The test was extended to 25% over the life target at which time parts were approaching the maximum lash change target value. 
     The valvetrain dynamics, Torsion spring load loss, pump-up validation and mechanical lash over an equivalent engine life all met intended targets for the SRFF-1L. The valvetrain dynamics, in terms of closing velocity, is safely within the limit at maximum engine speed of 7200 rpm and at the limit for a higher speed of 7500 rpm. The LMS load loss showed a maximum loss of 8% which is safely within the design target of 10%. A pump-up test was performed showing that the SRFF-1L design operates properly given a target oil pressure of 5 bar. Finally, the mechanical lash variation over an equivalent engine lift is safely within the design target. The SRFF-1L meets all design requirements for cylinder deactivation on a gasoline passenger car application. 
     6. Conclusions 
     Cylinder deactivation is a proven method to improve fuel economy for passenger car gasoline vehicles. The design, development, and validation of a single-lobe SRFF based cylinder deactivation system was completed, providing the ability to improve fuel economy by reducing the pumping losses and operating a portion of the engine cylinders at higher combustion efficiencies. The system preserves the base architecture of a standard Type II valvetrain by maintaining the same centerlines for the engine valves, camshaft and lash adjusters. The engine cylinder head requires the addition of the OCV and oil control ports in the cylinder head to allow for hydraulic switching of the SRFF from normal lift mode to deactivation mode. The system requires one OCV per engine cylinder, and is typically configured with four identical SRFF&#39;s for the intake and exhaust, along with one DFHLA per SRFF. 
     The SRFF-1L design provides a solution that reduces system complexity and cost. The most important enabling technology for the SRFF-1L design is the modification to the lost motion torsion spring. The LMS was designed to maintain continuous contact between a single lobe camshaft and the SRFF during both normal-lift and no-lift modes. Although this torsion spring requires slightly more packaging space, the overall system becomes less complex with the elimination of a three lobe camshaft. The axial stack up of the SRFF-1L is reduced from a three-lobe CDA design since there are no outer camshaft lobes that increase the chance of edge loading on the outer arm sliding pads and interference with the inner arm. Rocker arm stiffness levels for the SRFF-1L are comparable with standard production rocker arms. 
     The moment of inertia was minimized by placing the heavier components over the end pivot that sits directly on the DFHLA, namely the latching mechanism and the torsion springs. This feature enables better valvetrain dynamics by minimizing the effective mass over the valve. The system was designed and validated to engine speeds of 7200 rpm during standard lift mode and 3500 rpm for cylinder deactivation mode. The components also were validated to at least one engine life that is equivalent to 200,000 engine miles. 
     III. VVA Engine and Cylinder Head Arrangements 
     1. Switching Rocker Arm Assemblies 
     1.1. Description—General Engine Structure 
       FIGS. 115 and 116  illustrate a partial engine head assembly that is a conventional Type II, dual overhead cam internal combustion engine with the exhaust cam. Exhaust cam rockers, valves and a portion of the intake valve camshaft are removed for clarity. It should be noted here that the present teachings are equally applicable to other engine designs having similar arrangements and obstructions. 
     A plurality of cam towers  10  extend upward having a cam tower bottom  13  section that extends upward from the cylinder head. The upper side of the cam tower bottom  13  has a semi-circular recess. 
     A cam tower cap  11  is bolted to the cam tower bottom  13 . The cam tower cap  11  has a similar semi-circular recess facing downward such that when the cam tower cap  11  is bolted to the cam tower bottom  13 , the recesses create a circular cam recess  321  that receives the camshafts. Cam recesses  321  are sized and constructed to secure but allow the intake and exhaust camshafts to freely rotate. 
     Spark plug tubes  20  in this aspect of the present teachings are located between the cam towers  10  and parallel to a centerline  21  passing through the center of the cylinder head. The spark plug tubes  20  extend downward through the cylinder head into the top of each engine cylinder and are designed to receive a spark plug. 
     1.2. VVA Switching Rocker Arm Arrangements 
     1.2.1. Symmetrical Arrangements 
     This engine head assembly shown in  FIGS. 115 and 116  has enough space to accept a symmetrical variable valve lift (VVL) rocker arm assembly  100  as previously described. 
     The VVL rocker arm assembly  100  will be used for the remainder of the description provided here; however, it is understood that these aspects of the present teachings may be applied to various other rocker arm assemblies installed on heads having small clearances on one side of the rocker assemblies. 
     This VVL rocker arm assembly  100  is driven by a camshaft having three lobes per cylinder. It is shown here in  FIGS. 115 and 116  with the camshafts removed except a middle cam lobe  324  and an outer cam lobe  326  remain and are shown. In this aspect of the present teachings, a rocker arm assembly  100  is shown that has an inboard end  101  (or a first end  101 ) and an outboard end  103  (or a second end  103 ). The term ‘inboard’ refers to a direction inward toward centerline  21  and ‘outboard’ refers to a direction outward away from the centerline  21 . 
     As seen in  FIG. 116 , it is seen that the VVL rocker arm assembly  100  inboard end  101  is supported by a hydraulic lash adjuster  340 . The outboard end  103  rests upon valve stem  350 . 
     As middle cam lobe  324  turns and presses downward onto the VVL rocker arm assembly  100 , it causes outboard end  103  of VVL rocker arm assembly  100  to push valve stem  350  downward opening the poppet valve connected to valve stem  350 . When an internal latch is operated by providing high-pressure oil to it, the VVL rocker arm assembly  100  causes the valves to lift according to the shape of the outer cam lobes  326 . This is further described below in connection with  FIG. 117 . 
     1.2.2 Non-Symmetrical Arrangements 
     In  FIG. 117 , the torsion springs  135 ,  137  and spring posts  141 ,  143  make the VVL rocker arm assembly  100  wider at its first end as compared with a standard rocker arm. The design of the VVL rocker arm assembly  100  (and that of the CDA rocker arm) is wider than standard rocker arms and can fit in only certain cylinder heads. There is enough clearance in the cylinder heads shown in  FIGS. 115 and 116 , however, in certain engine heads, there is not enough clearance from other structures, such as a cam tower or spark plug tube, and this VVL rocker arm assembly  100  could not be used. 
     As indicated above, it is very costly to redesign/modify cylinder heads, cam drives and gear trains. Also, many different manufacturers may make equipment based upon the standard design of the cylinder head, making it very difficult to change or modify the cylinder head. 
     Therefore, the present teachings can be embodied in VVA rocker arm assemblies that are specially designed to fit cylinder heads having little clearance. 
     In many cylinder head designs, it was determined that there was only a lack of space in one side of the rocker. Typically, the lack of space can occur in the inboard end  101  on the side of the rocker near the spark plug tubes  20 . Therefore, it would be viable to package the VVL rocker arm assembly  100  in a redesigned form so that the width on the obstructed side would not be wider than that of a standard rocker arm. 
     The result was to create modified rocker assemblies for use on cylinder heads that have obstructions on the right-hand side of the rocker assemblies, or left-hand rocker assemblies. In the left-hand rocker assembly most of the functional elements are moved from the right-hand side to the left-hand side. Also, the right-hand side is formed to have reduced width. 
     Similarly, right-hand rocker assemblies are designed for use when there is an obstruction on the left-hand side. Similarly, structures are moved from the left-hand side to the right-hand side and the left-hand side is formed to create increased clearance on the left side to compensate for the obstruction. Collectively, they will be referred to as modified rocker assemblies. 
     A novel modified rocker assembly  400  according to one aspect of the present teachings is described in connection with  FIGS. 118-122 . 
       FIG. 118  is a perspective view of a left-handed modified rocker assembly  400  exhibiting variable valve lift, according to one aspect of the present teachings. 
       FIG. 119  is top plan view of the modified rocker assembly  400  of  FIG. 118 . 
       FIG. 120  is a side elevational view of the modified rocker assembly  400  of  FIGS. 118-119 . 
       FIG. 121  is an end-on elevational view of the modified rocker assembly  400  of  FIGS. 118-120  as viewed from its hinge (first) end. 
       FIG. 122  is an end-on elevational view of the modified rocker assembly  400  of  FIGS. 118-121  as viewed from its latch (second) end. 
     The modified rocker assembly  400  shown here for illustrative purposes is a variable valve lift (VVL) rocker assembly; however, a cylinder deactivation (CDA) rocker assembly or other rocker assembly employing torsion springs on its first end  408 , or otherwise having a widened first (or hinged) end  408  fall within the scope of the present teachings. 
     This rocker assembly functions in a very similar manner as that shown in  FIG. 117  and described above, and the VVL Rocker Application, herein incorporated by reference. The modified rocker assembly  400  employs an inner structure  410  that fits inside of an outer structure  420 . However, this modified rocker assembly  400  is used on cylinder heads having less clearance near the rocker assembly. The modified rocker assembly  400  includes many ornamental aspects apart from the functional aspects disclosed herein. 
     Inner structure  410  can have an axle recess  413  passing through its first end  408 . The outer structure  420  also can have an axle recess  433  through its first end  408 . When the roller axle recesses  413 ,  433  are aligned with the inner structure  410  inside of the outer structure  420 , the axle  434  can be secured through the axle recesses  413 ,  433  allowing inner structure  410  to pivot relative to outer structure  420  about axle  434 . 
     The outer structure  420  on the obstructed side  405 , as it extends from the second end  409  toward the first end  408 , can be offset toward the non-obstructed side  407  creating a first offset portion  428 . This offset can be a curved or angled sidearm that can create a smaller width at the first end  408 . This first offset portion  428  can provide additional clearance on the obstructed side  405  as compared with standard VVL or CDA rocker arm assemblies. This can now allow the modified rocker assembly  400  to fit into and function with cylinder heads that have narrow obstruction region such as obstruction region  600  of  FIGS. 132, 133 . 
     The outer structure  420  on the non-obstructed side  407 , as it extends from the second end  409  toward the first end  408 , can be offset outward away from the modified rocker assembly  400  creating a second offset portion  429 . This second offset portion  429  can provide additional clearance on the non-obstructed side  407  as compared with standard VVL or CDA rocker arm assemblies, to allow the incorporation of a second torsion spring  437 . This now can allow the modified rocker assembly  400  to exert the proper amount of force to bias the inner structure  410  with respect to the outer structure  420 . In an alternative aspect of the present teachings, a single larger torsion spring can be used in place of the two or more torsion springs shown here. 
     The modified rocker assembly  400  employs a latch assembly  500  with a latch pin  501  that can hold the inner structure  410  and outer structure  420  together so they move as a single rocker. The latch assembly  500  can be activated by an oil control valve (not shown) that can provide increased oil pressure through a cup  448  pivoting upon the hydraulic latch adjuster  340 . This is further described in connection with  FIGS. 126, 127 . 
     Since there are now two (or more) torsion springs  435 ,  437  on the non-obstructed side  407  (or here is a single larger torsion spring) with no torsion springs on the obstructed side  405 , there will be a twisting force placed upon the inner structure  410  and outer structure  420  of the rocker assemblies. Therefore the amount of play about the axle  434  can be adjusted to make sure that the modified rocker arm  400  functions correctly. 
     When using two torsion springs  435 ,  437 , torsion spring  435  is considered a right-hand side spring and is wound in the opposite direction of torsion spring  437 . These different springs null out some of the torsional forces. 
     If only a single torsion spring is to be used, the additional torsional forces should be considered when designing the inner and outer structures  410 ,  420 . 
     For the double torsion spring and single torsion spring designs, the relative strength of the inner and outer structures  410 ,  420  can be adjusted to reduce flexing, to ensure proper performance. Also the weight distribution of each of the structures along their length can be configured to provide the proper strength and structure while minimizing the inertial force required to pivot the modified rocker assembly  400  at the speed required to operate an engine. 
       FIG. 122  shows the latch pin seat  485  that receives and holds latch pin  501  when it is in the extended position. Latch pin  501  and latch pin seat  485  can hold inner structure  410  from fitting into outer structure  420 . Even though the latch pin is shown as a round shape, it may have a flat end that corresponds to a flat seat. The latch pin  501  and latch pin seat  485  can have any complementary shape that allows them to fit properly together. 
       FIG. 123  is a plan view from above the outer structure  420  showing the first and second offset areas  428 ,  429 . Here the differences from the outer structure of the rocker assembly of  FIG. 117  can be seen. The first outer side arm  421  near the first end  408  can be skewed to the left to accommodate an obstruction on the right side of the first end of rocker assembly  400 . Similarly, the second outer side arm  430  can also be skewed to the left to also accommodate an obstruction on the right side of the first end of rocker assembly  400 , keeping the first and second outer side arms roughly the same distance from each other as they extend from the second end  409  toward the first end  408 . This can create the offset areas  428  and  429 . 
       FIG. 124  is a plan view from below the outer structure  420  of  FIG. 123  also showing the first and second offset areas  428 ,  429 . This also shows a lower cross arm  439 . The lower cross arm  439  can be shown to add strength to counteract forces and help prevent flexing that may otherwise occur, due to the non-symmetric design of the modified rocker assembly  400 . 
     Latch pin seat  485 , discussed in connection with  FIG. 122  above, is also visible from this view. 
       FIG. 125  is a side elevational view of an outer structure  420  according to one aspect of the present teachings. The first outer side arm  421  and first offset portion  428  are visible in this view. 
       FIG. 126  is a perspective view of top side of an inner structure  410  according to one aspect of the present teachings. This is a view from the obstructed side showing axle recess  413 . This is designed to have the spring post  447  on the non-obstructed side. This design allows additional space on the obstructed side. 
       FIG. 127  is a perspective view of bottom side of the inner structure  410  of  FIG. 126 . Axle recess  413  is shown that can receive axle  434  and can pivotally connect the inner structure  410  to the outer structure  420 . In both  FIGS. 126 and 127 , roller axle apertures  483  and  484  can receive the roller axle (not shown) to hold roller  415 . In  FIG. 127 , cup  448  can receive the hydraulic lash adjuster  340  of  FIG. 116 . The hydraulic lash adjuster ( 340  of  FIG. 116 ) is provided with oil flow from an oil control valve (not shown). The hydraulic lash adjuster  340  has an oil outlet that can provide the oil flow into cup  448 . Cup  448  can be connected to internal passageways that provide the oil to galleries  444  and  446 . The oil galleries can be connected by internal passageways to latch assembly  500 . An oil pressure provided by the oil control valve greater than a threshold pressure can cause the latch assembly  500  to be switched. The latch pin ( 501  of  FIGS. 120-122 ) can be set to its normal position (with low oil pressure) in a retracted position. When the oil pressure greater than a threshold amount is provided to the latch, it can switch to extend latch pin ( 501  of  FIGS. 120-122 ). This is a ‘normally unlatched’ configuration. 
     Alternatively, at low oil pressure, the latch pin can normally be in an extended position. When the oil pressure increases above a threshold amount, the latch pin can be retracted. This is a ‘normally latched’ design. 
       FIG. 128  is a plan view from the top side of the inner structure of  FIGS. 126-127 . 
       FIG. 129  is a plan view from the bottom side of the inner structure of  FIGS. 126-128 . 
     In  FIG. 129 , the valve stem seat  417  is shown. Valve stem seat  417  presses against the engine valve stem, actuating the valve when the modified rocker assembly  400  pivots. 
       FIG. 130  is an end-on elevational view of the inner structure  410  of  FIGS. 126-129  as viewed from its hinge (first) end. 
       FIG. 131  is an end-on elevational view of the inner structure  410  of  FIGS. 126-130  as viewed from its latch (second) end. 
     In  FIGS. 128-131  spring post  447  is shown. One or more of the first torsion springs  435 ,  437  fit over and can be held in place by the spring post  447 . A single larger torsion spring may also be used in place of first and second torsion springs  435 ,  437 . 
       FIG. 132  is a perspective view of the modified rocker assembly  400  of  FIGS. 118-122  as it would appear installed in a cylinder head. 
     As with  FIGS. 115 and 116 , parts have been removed for clarity. Most notably, the shaft portion of a camshaft having three lobes per engine valve has been removed. The middle cam lobe  324  and one outer cam lobe  326  are shown. Since one of the side lobes is not shown, the second slider pad  431  is visible. This and the second slider pad can ride on the outer cam lobes  326  as described in the VVL Rocker Application above. 
     The camshaft would be secured by and pass through the cam tower  10 . Here it can be easily seen that spark plug tube  20  would interfere with a standard CDA or VVL rocker assembly at the obstruction region  600 . The first offset portion  428  of the modified rocker assembly  400  is adjacent to the spark plug tube  20  at obstruction region  600 . Due to its reduced width, it is now able to fit on this head and function without colliding into the spark plug tube  20 . 
       FIG. 133  is a perspective view from another viewpoint of the modified rocker assembly  400  of  FIGS. 118-122 , as it would appear installed in a cylinder head. 
     This shows the same structures as  FIG. 120 , but from a viewpoint above, and closer to the centerline of the cylinder head, viewing the non-obstructed side  407  of the modified rocker assembly  400 . Middle cam lobe  324  is pressing down roller  415 . 
     First offset portion  428  is shown near the obstruction region  600  adjacent to the spark plug tube  20  providing the required clearance. 
     Second offset portion  429  is also shown providing the additional space for both torsion springs  435 ,  437 . 
     While the present disclosure illustrates various aspects of the present teachings, and while these aspects have been described in some detail, it is not the intention of the applicant to restrict or in any way limit the scope of the claimed teachings of the present application to such detail. Additional advantages and modifications will readily appear to those skilled in the art. Therefore, the teachings of the present application, in its broader aspects, are not limited to the specific details and illustrative examples shown and described. Accordingly, departures may be made from such details without departing from the spirit or scope of the applicant&#39;s claimed teachings of the present application. Moreover, the foregoing aspects are illustrative, and no single feature or element is essential to all possible combinations that may be claimed in this or a later application.