Abstract:
Improved liquid lubricated piston ring gas seals with reduced friction and wear compared to prior art rings are disclosed for use in internal combustion engines, gas pumps, and gas compressors. These improved rings allow independent adjustment of the parameters controlling friction and wear, and eliminate the tradeoff between friction and wear typical of prior art rings.

Description:
This application is a division claiming priority from U.S. application Ser. No. 14/428,072, filed Mar. 13, 2015. 
     FIELD OF THE INVENTION 
     The present invention is directed to piston rings that form sliding seals in pistons operating in cylindrical bores, and more particularly relates to liquid lubricated compression rings used as piston gas seals in internal combustion engines and reciprocating pumps and compressors. 
     BACKGROUND OF THE INVENTION 
     The present invention comprises improvements to the reduced friction piston rings described in Meacham WO/2009/033115, which is incorporated by reference in its entirety. 
     It is generally known to provide liquid lubricated piston ring seals in reciprocating pistons to reduce gas flow through the diametral clearance between the piston and the bore in which it reciprocates. 
     Compression rings of the type typically used in internal combustion engines are heat and wear-resistant hard materials such as metal, and may have metallic or ceramic coatings to improve their friction and wear properties. Compression rings are generally circular with a rectangular cross section and a small radial gap and are installed in annular grooves in the pistons. Prior to installation the outside diameter of the rings is slightly larger than the inside diameter of the bores. The radial gap allows the rings to be elastically expanded so that they can be installed in the piston grooves. When the pistons are installed in the bores, the rings are elastically compressed to the smaller bore diameter such that the radial gaps are nearly closed. The bore, the cylindrical piston sides, the rings, and the grooves are coated with a thin layer of liquid lubricant, e.g. mineral oil. In the installed condition the rings exert a moderate radial elastic force against the bore surfaces to provide a baseline sealing force. When pressure is applied to a piston and ring assembly, the pressure difference presses the rings against the low-pressure sides of the grooves, and pressurized gas flows into the ring grooves between the piston and the rings. This pressurized gas exerts outward radial force on the rings that augments the elastic baseline force. 
     The value of the pressure-driven outward radial force is directly proportional to the differential pressure across the ring. The inward balancing force depends on the details of the contact interface acting on a portion of the contact interface areas between the ring and the bore surface. The unbalanced portion of the force is carried as a sliding bearing contact load between the ring and the cylinder bore. At high piston speeds typical of most of the stroke, the ring-bore sliding bearing is supported hydrodynamically on a liquid lubricant film without metal-to-metal contact. Boundary lubrication with metal-to-metal contact occurs as the piston slows and reverses at the stroke ends. This results in much higher friction, and causes most of the ring wear. The simplest case is uniform contact across the entire axial ring thickness. The contact interface area pressure near the edge exposed to the high pressure gas is equal to the high pressure, and the contact interface area pressure near the other edge exposed to the low pressure gas is equal to the low pressure. The pressure at any point between the edges is intermediate between these values. The total inward radial pressure force is less than the outward radial pressure force, and the resultant ring sealing force therefore increases with increasing differential pressure. If the pressure variation between the edges is linear, the inward radial pressure force balances about half the outward radial pressure force. Small changes in variables such as ring twist can make substantial changes in the pressure variation between the edges. Twist that opens a gap towards the high pressure edge increases the total inward radial pressure force. This reduces friction, but may increase blow-by. Twist that opens a gap towards the low pressure edge decreases the total inward radial pressure force. This increases friction and may reduce blow-by. The net outward force on the ring is the vector sum of the outward radial pressure force, the inward radial pressure force, the outward radial ring elastic force and the radial friction force between the sides of the ring and the piston ring groove. This net outward force and the friction coefficient determine the friction between the ring and the bore. If also determines the load supported by the sliding bearing formed by the contact zone between the ring outside diameter and the cylinder bore. The ring friction and wear are critically dependent on the presence of liquid lubricant, the contact zone area, the piston velocity and the net outward force. Conventional piston rings represent a difficult compromise between design parameters to achieve the best possible friction and wear performance for a given application, and a principal feature of this invention is a reduced need to compromise. 
     Since friction increases with increasing total outward ring radial force, one piston ring design objective is to reduce this force as much as possible consistent with keeping the ring loaded against the bore to maintain a gas seal. The most direct way to reduce the outward ring sealing force of conventional rings is to reduce the ring thickness in the axial direction. This limits the maximum outward pressure force regardless of magnitude of inward balancing force. There are limits, however, on how thin the rings can be made and survive in the engine environment. Further, the top compression ring forms a significant thermal conduction path between the piston and the cooler cylinder bore that is important in keeping the piston cool, and narrow rings reduce the thermal conduction. Consequently, many engine designs have relatively thick robust rings configured such that inward radial pressure force balances a large part of the outward radial pressure force. A variety of ring cross section contours and twist conditions have been proposed to do this. A convex ring barrel bore contact surface with line contact defines the interface areas exposed to high pressure and low pressure precisely. Alternatively a tapered ring outer bore contact surface defines a substantial outer ring area exposed to high pressure during the compression and power strokes. The effective size and position of contact areas defined by convex barrel shapes or tapers are, however, affected by relatively small amounts of ring wear. One solution is application of a hard coating to limit ring wear so that the geometry is maintained during ring break-in and service. Another is to use a ring cross section in which the high pressure interface area is largely defined by a raised flange on the outer bore contact surface, resulting in a design with controlled radial force that is insensitive to ring wear. 
     The liquid lubricant film has an important effect on piston ring performance. During mid-stroke the high piston velocity causes the liquid lubricant to form a hydrodynamic film between the ring outer bore contact surface and the cylinder bore that prevents metal to metal contact. In this hydrodynamic mode the ring acts as a linear slider bearing of the type described in the Standard Handbook for Mechanical Engineers, 7th Edition, edited by Theodore Baumeister. Page 8-171 (1960) published by the McGraw-Hill Book Company, New York. This lubrication theory indicates that the hydrodynamic coefficient of friction f h  is: 
             f     h   ⁢           ∝       μγC   W               
where μ is lubricant viscosity, γ is piston velocity, C is the bore circumference and W is the radial load supported. F h  is generally low. The coefficient of friction f h  in boundary lubrication increases dramatically because of metal to metal contact between the ring and the bore. Significantly the hydrodynamic coefficient of friction is independent of axial ring bearing contact width. This theory also indicates that the lubricant film thickness h o  is:
 
             h     o   ∝     L   ⁢       μγC   W                 
where L is the axial ring bearing contact width resulting in increased film thickness with increased contact width. The conclusion is that increased axial ring bearing contact width increases the lubricant film thickness at a given piston velocity and lubricant viscosity without increasing the friction coefficient.
 
     As the piston slows and reverses at the stroke ends, the lubricant film thickness h o  decreases, and hydrodynamic lubrication transitions to boundary lubrication with metal to metal contact and increased friction and wear. Increased bearing contact width increases lubrication film thickness h o  and delays the transition to boundary lubrication. Alternatively, increased bearing contact width might be used advantageously to reduce lubricant viscosity rather than delaying the transition to boundary lubrication, allowing reduced friction elsewhere in the engine that more than offsets the increased boundary layer friction of the piston rings. 
     The transition to boundary lubrication may also be delayed by squeeze film lubrication. Squeeze film lubrication is a transient process in which the lubricant film is squeezed out of the gap between the ring and the cylinder bore as the surfaces approach metal-to-metal contact. The film squeezing process requires a period of time that increases with the initial gap, lubricant viscosity and contact width. This time period extends the effective hydrodynamic lubrication regime and reduces the boundary lubrication regime. 
     In theory increased contact width maximizes the hydrodynamic bearing regime and minimizes the boundary lubrication regime without changing the friction coefficient at constant lubricant viscosity and radial ring force. The question is how to utilize these effects. Conventional rings with increased contact width, as discussed earlier, tend to have high outward radial pressure force and high friction that negates improved bearing performance. Conventional rings therefore typically compromise towards low contact width to minimize radial force and friction through most of the stroke and accept a certain amount of boundary lubrication sliding and wear at the stroke ends. The objective of this invention is to provide piston ring configurations that effectively increase the bearing area while minimizing the radial ring force and friction, and enhance the squeeze film effect to reduce friction and wear at the stroke ends. 
     SUMMARY OF THE INVENTION 
     The present invention is directed to designs and methods for reducing friction and wear of liquid lubricated compression piston ring gas seals. Piston rings according to the invention combine low but well defined outward radial pressure force with increased ring area contacting the bore. The low outward radial pressure force reduces friction throughout the piston motion. The increased ring area contacting the bore extends the hydrodynamic and squeeze film lubrication regime through a larger portion of the piston motion to minimize wear and further reduce friction. Alternatively, the increased ring area contacting the bore may allow reduced liquid lubricant viscosity that reduces friction elsewhere in the engine. The description focuses on compression piston rings for internal combustion engines, but the present invention is applicable to piston rings used in reciprocating piston pumps, gas compressors and other applications using liquid lubricated piston ring gas seals. 
     According to the invention, the compression piston ring is configured such that the contact interface area between the ring and the bore surface is divided into a seal zone and bearing zones. The seal zone separates high pressure gas on one side of the piston ring from low pressure gas on the opposite side. As with conventional rings, the average gas pressure in the seal zone is intermediate between the high pressure and the low pressure, and generates less inward radial force than if the total seal zone area was exposed to the high gas pressure. The bearing zones are also in contact with the bore, but do not separate high pressure gas from low pressure gas. Instead, each bearing zone is entirely surrounded by gas at a balancing pressure that is approximately the same pressure as the gas under the ring that generates the outward radial pressure force. It therefore generates an inward radial force that counteracts or balances a portion of the outward radial force. The invention includes multiple ways to form seal zones and bearing zones and means of maintaining the gas sealing and bearing functions as the piston tilts in the bore. 
     The bearing zones are used to control the ring radial pressure force balance as an improved alternative to conventional design features that form annular gaps between the ring and bore to admit gas at the balancing pressure to reduce friction. These conventional design features include chamfers, tapers, convex barrel contact surfaces, ring twist, and recessed faces. Compared to these conventional design features, the use of bearing zones provides more ring bearing area sliding against the bore. This brings several advantages. The bearing zone area supplements the seal zone bearing area, but without increasing the outward radial force. This allows a small seal zone area with low net outward radial force and ring friction, while the larger total bearing area operates at reduced contact pressure and protects the ring, including the seal zone, from excessive wear. Moreover, the fact that the seal zone and bearing zones are defined by distinct geometric features, rather than subtle features such as taper or ring twist, allows consistent performance over a large wear range. Further, the increased contact area results in a thicker hydrodynamic liquid lubricant film and larger squeeze film lubrication effect that reduces friction and wear at the stroke ends. Finally, the increased contact area also enhances heat transfer from the ring and piston to the cylinder bore, thereby reducing thermal stress on the piston, rings and lubricant. Altogether, these factors may reduce wear and minimize the need for hard wear-resistant coatings. 
    
    
     
       DESCRIPTION OF DRAWINGS 
       The appended claims set forth those novel features that characterize the invention. However, the invention itself, as well as further objects and advantages thereof, will best be understood by reference to the following detailed description of preferred embodiments. The accompanying drawings, where like reference characters identify like elements throughout the various figures in which: 
         FIG. 1  illustrates a one-piece reduced friction ring in an internal combustion engine: 
         FIG. 2  provides perspective views of the one-piece reduced friction piston ring: 
         FIG. 3  illustrates the function of the one-piece ring as the piston tilts in the bore: 
         FIG. 4  illustrates a two-piece reduced friction ring in an internal combustion engine: 
         FIG. 5  provides perspective views of the two-piece reduced friction piston ring: 
         FIG. 6  illustrates the function of the two-piece ring as the piston tilts in the bore: 
         FIG. 7  illustrates an articulated two-piece reduced friction ring in an internal combustion engine: 
         FIG. 8  provides perspective views of the articulated two-piece reduced friction piston ring: and 
         FIG. 9  illustrates the function of the articulated two-piece ring as the piston tilts in the bore. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     Upon examination of the following detailed description the novel features of the present invention will become apparent to those of ordinary skill in the art or can be learned by practice of the present invention. It should be understood that the detailed description of the invention and the specific examples presented, while indicating certain embodiments of the present invention, are provided for illustration purposes only. Various changes and modifications within the spirit and scope of the invention will become apparent to those of ordinary skill in the art upon examination of the following detailed description of the invention and claims that follow. 
     The prior art and the invention are described with reference to internal combustion engines, but it is to be understood that the invention is applicable to liquid lubricated piston ring gas seals in other applications including gas compressors. In the description “upper”, “top”, “above” and “head” refer to the direction towards the combustion chamber, and “lower” and “downward” refer to the direction towards the crankcase. 
     The description focuses on the effects of pressure and friction forces and piston tilt on the rings. It is recognized that inertial forces are also present, but these are not central to the invention and are included in the description only in relation to ring lift. Further the description is limited to the compression and expansion strokes wherein high gas pressures occur above liquid lubricated compression piston rings, generating high pressure-driven radial contact forces and resulting ring friction and wear. Intake and exhaust strokes are not discussed because the low pressures generate only small pressure-driven radial contact forces. 
       FIG. 1 ,  FIG. 2  and  FIG. 3  show a first improved version of a one-piece reduced friction pressure balanced top compression ring described in detail in Meacham WO/2009/033115. The ring  100  is installed in the top groove  105  of the piston  110  such that it slides against the cylinder bore  108  and forms a seal blocking high pressure gas  109  from flowing through the clearance between the piston  110  and the bore  108 . The piston typically also employs a second compression ring  111  and an oil control ring  112  in lower grooves. As with the original version, the new ring  100  has a seal zone  101  and a bearing zone  102  on the ring outside diameter separated by a circumferential groove  103 , where this groove is connected to the top of the ring by axial grooves  104  crossing the bearing zone. The axial grooves allow gas  109  at cylinder pressure to flow to the circumferential groove  103 , so groove  103  is at approximately the same pressure as the piston groove  105  and the inside diameter  106  of the ring, which are also exposed to gas at cylinder pressure. 
     As pressure rises in the cylinder, ring  100  is pushed down against the lower flank  116  of the piston groove  105 . The clearance between the upper piston groove flank  107  and the top ring surface  114  allows gas  109  to flow into the piston groove and pressurize the inside diameter  106  of the ring, creating an outward radial force pressing the ring  100  against the cylinder bore  108 . Since the axial grooves  104  cause the circumferential groove  103  pressure to be approximately equal to the pressure in the ring groove  105 , the interface pressure between the bearing zone  102  outside diameter and the cylinder bore  108  is approximately equal to the cylinder pressure  109  contacting the ring inner diameter  106 . The effect of this interface pressure is to generate an inward radial force that partially balances the outward radial force on the ring, with the result that the bearing zone  102  does not contribute to the pressure driven outward radial force and friction of the ring. The only radial pressure force is generated by the seal zone  101 , and this is low when the seal zone is narrow. The bearing zone  102 , however, carries a portion of the radial force generated in the seal zone  101 . Since the bearing zone  102  is relatively wide, it maintains hydrodynamic lubrication over a larger part of the engine cycle and reduces friction and wear. It also improves heat transfer between the ring and piston assembly and the cylinder bore. 
     One improvement in the design shown in  FIG. 1  and  FIG. 2  is an optional array of radial vent grooves  113  in the top face of the ring to increase the flow rate of gas into the ring groove to more closely match the pressure rise rate in the bearing zone without increasing the mechanical clearance of the ring in the groove. This is advantageous since it minimizes the possibility of transient ring collapse and seal leakage caused by a faster pressure rise in the bearing zone  102  than the inside ring diameter  106 . 
     The second improvement is a contour on the outside diameter  115  of the ring  100  that optimizes sealing performance as the piston  110  tilts in the bore  108 . The seal zone  101  has a barrel shaped contour, and the bearing zone  102  has a tapered contour where the taper half-angle is equal to or greater than the maximum piston tilt. The taper and barrel contours are dimensioned such that if the taper were extended down across the circumferential groove  103  it would be tangent to the upper edge of the seal zone barrel contour. This geometry allows the seal zone  101  to remain in contact with the bore  108  as the piston  110  tilts. It should be noted that the taper and barrel contours shown in the illustrations are exaggerated to clarify the principle. The actual taper is on the order of 1 to 2 microns per millimeter of bearing zone width. 
     The arrangement provides a seal ring with well-defined bearing and seal zones that can be separately optimized. The sealing zone may be very narrow to reduce the outward radial pressure force and friction, while a wide bearing zone can be used to increase the hydrodynamic bearing oil film thickness and reduce metal-to-metal contact and resulting friction and wear. 
     The combination of a barrel shape in the sealing zone  101  and a tangential taper in the bearing zone  102  maintains a seal under tilt conditions as illustrated in  FIG. 3 . The minimum clearance between the ring  100  and the cylinder bore  108  is in the seal zone  101 , and clearance increases by a few microns in the tapered bearing zone  102 . This prevents contact between the bearing zone  102  and the cylinder bore  108  from lifting the seal zone from the cylinder bore and increasing blow-by as the piston  110  tilts. The taper is small enough that the bearing zone is still effective as a hydrodynamic slider bearing. It does, however, limit the ability of the bearing zone  102  to protect the seal zone  101  from wear in boundary lubrication conditions. 
     Cylindrical seal and bearing zones that allow a small amount of seal zone lift with piston tilt are an alternative that maximizes the bearing performance and wear resistance of the ring assembly at the expense of increased blow-by. The blow-by increase is expected to be small. A one micron gap on one side of a 75 millimeter bore piston forms a flow area of 0.12 square millimeters, equivalent to a 0.4 millimeter round hole. This is a small leak that only occurs during piston tilt, so it may a good trade-off for some applications. 
     Manufacturing tolerances are optimized for this one-piece ring configuration so that the seal zone  101  outside diameter extends radially beyond the bearing zone  102  outside diameter at the initial assembly of the piston  110  in the bore  108  to assure good sealing. The seal zone  101  will wear slightly during engine break-in and allow the bearing zone  102  to share more of the radial load. 
     In addition to carrying part of the seal zone  101  radial pressure load to reduce wear, the bearing zone  102  shields the narrow seal zone from direct contact with the high temperature combustion gas  109 . Instead, the gas flows through the small axial grooves  104  across the bearing zone  102  and is cooled before contacting the seal zone  101 . This shielding is expected to improve the seal zone durability since it reduces its operating temperature. The bearing zone  102  also contributes to heat transfer from the piston  110 , since it has a relatively large area in contact with the piston bore. 
       FIG. 4 ,  FIG. 5  and  FIG. 6  show a second improved version of the one-piece reduced friction pressure balanced top compression ring described in detail in Meacham WO/2009/033115. Installation and function of this two-piece ring assembly is the same in many respects as the one-piece ring  100  described with respect to  FIG. 1 ,  FIG. 2  and  FIG. 3 , and only differences will be described in detail. The ring is an assembly of a seal ring  400  and a bearing ring  401 . The two rings interlock via a circumferential groove  402  on an inner face  404  of the seal ring  400  and a mating lug  403  on the adjacent face  405  of the bearing ring  401  so that the two rings act as a single unit when assembled into the piston groove  105 . Radial pressure load is thereby transferred from the seal ring  400  to the bearing ring  401 . A chamfer  406  on the upper outside diameter of the seal ring forms a circumferential volume  410  between the seal ring and the bearing ring. Radial vent grooves  407  in the bearing ring  401  provide gas flow passages from the ring inner diameter  408  to the outer diameter  409  so that the circumferential volume  410  formed by the seal ring chamfer is held at a value approximately equal to the pressure in the piston groove  105  acting on the ring inner diameter  408 . The circumferential volume  410  has the same function as the circumferential groove  103  in the first ring  100 , and the radial vent grooves have a similar function as the axial vent grooves  104 . The principal difference is that the gas providing the balancing pressure in the circumferential volume  410  flows from the piston groove  105 , and the pressure rise in this volume will lag the rise in cylinder pressure  109 . This in turn assures that the pressure rise in the bearing zone  401  lags the pressure rise in the piston groove  105 , which is advantageous since it minimizes the possibility of transient ring collapse and seal leakage caused by a faster pressure rise in the bearing zone  401  than in the piston groove  105 . 
     The lug  403  and groove  402  are positioned at the neutral axis of each ring to minimize ring twist effects when the rings are elastically bent to provide the baseline radial force against the bore surface. It also minimizes stress concentrations or ring distortions caused by the periodic radial vent grooves  407  in the seal ring  400  circumferential lug  403 . If a twist bias is needed, it may be added in the conventional manner by chamfering the ring inside diameter. 
     The contours on the outside diameters of the seal ring  400  and bearing ring  401  combine to provide the same function as the contours of the one-piece ring  100  to optimize sealing performance as the piston tilts in the bore. The seal ring  400  has a barrel shaped contour, and the bearing ring  401  has a tapered contour. Cylindrical seal and bearing zones that allow a small amount of seal zone lift with piston tilt are also an alternative to maximize the bearing performance and wear resistance of the ring assembly at the expense of increased blow-by. In this case the initial dimensions of the rings should favor contact between the seal ring  400  and the bore  108 , allowing break-in wear to bring the bearing ring  401  into full contact. The heat transfer and heat shielding characteristics of the one-piece and two-piece variations are also similar. 
     There are additional differences between the one-piece and two-piece variations. The seal ring  400  and the bearing ring  401  may be made of different materials to optimize the cost and performance of the assembly. The seal ring, for example, might be nitrided steel for high strength, while the bearing ring might be cast iron for good lubricity and low wear. Also, the two-piece ring assembly has reduced axial ring lift caused by inertial forces than a one-piece ring, since the seal ring  400  has lower mass. The bearing ring  401  may lift, leaving the seal ring in contact with the lower groove face. 
       FIG. 7 ,  FIG. 8  and  FIG. 9  show a third improved version of the one-piece reduced friction pressure balanced top compression ring described in detail in Meacham WO/2009/033115. Installation and function of this two-piece ring assembly is the same in many respects as the one-piece ring and the two-piece ring described with respect to  FIG. 1  through  FIG. 6 , and only differences will be described in detail. The two-piece articulated pressure balanced ring accommodates piston tilt by providing a separate seal ring  700  and bearing ring  701  that move relative to each other to optimize both bearing and sealing performance as the piston tilts. A circumferential lug  702  on the top side of the seal ring portion of the assembly engages a circumferential groove  703  in the bottom side of the bearing ring  701  to transfer radial pressure loading on the seal ring  700  to the bearing ring  701 . The lug  702  also acts as an axial spacer to form an axial gap  704  between the seal and bearing rings. This gap, combined with periodic radial notches  705  in the seal ring circumferential lug  702 , provides gas flow paths from the inner diameter  706  of the ring assembly to the lower edge of the bearing ring  701  outer diameter  707  to provide the pressure balancing function. Further, the lug and groove connection between the seal ring  700  and bearing ring  701  allows articulation of the assembly such that the seal ring  700  follows the piston tilt to maintain a tight seal and the bearing ring  701  rocks within the piston groove  105  to align with the bore  108  to carry radial force. An optional crown radius  710  on the upper face of the bearing ring  701  allows it to rock within the piston groove  105 , while maintaining minimal axial clearance of the ring assembly within the piston groove. 
     The lug  702  and groove  703  are shaped to provide pivot and cam functions that maintain an optimal geometric relationship between the two rings  700  and  701 , the piston  110  and the bore  108  over the range of piston tilt. A circumferential pivot  708  is formed in the top of the bearing ring groove  703 , and bears on the top of the seal ring lug  702  to allow the bearing ring  701  to rock relative to the seal ring  700  while defining the center of rotation. The outer diameter of the lug  702  and the outer diameter of the groove  703  are shaped such that they engage as a cam  708  and follower  709  to adjust the relative radial position of the seal ring  700  as the bearing ring  701  rocks. 
     This articulation action is shown in  FIG. 9 . Gas pressure keeps the seal ring  700  loaded against the lower piston groove face  106  as the piston tilts, and the large barrel radius on the outer diameter  706  maintains a good seal as the seal ring  700  and piston  110  tilt together. At the same time the radial spring force of the bearing ring  701  plus the radial pressure loading transferred from the seal ring  700  keeps the bearing ring aligned with the cylinder bore  108 . This requires the bearing ring  701  to rock within the groove  105 , and the crown  710  on the upper face of the bearing ring  701  allows this rocking motion. With counterclockwise rotation of the piston  110  the cam  708  and follower  709  move relative to one another to pull in the seal ring  700  relative to the bearing ring  701  to maintain the radial pressure load transfer to the bearing ring. With clockwise rotation of the piston  110  the cam  708  and follower  709  move relative to one another to allow the seal ring  700  relative to move out relative to the bearing ring  701  and maintain sealing contact with the cylinder bore  108 . It should be noted that the radial adjustments are in the opposite directions on the opposite sides of the tilted piston  110 . Since the seal ring  700  is aligned with the piston groove  105  and the bearing ring  701  is aligned with the bore  108 , neither ring has to twist as the piston tilts. 
     The heat transfer and heat shielding characteristics of the two-piece articulated pressure balanced ring assembly are similar to the one-piece and non-articulated two-piece ring variations. It shares the improved axial ring lift and radial collapse characteristic of the non-articulated two-piece ring variations, as well as the ability to manufacture the seal ring  700  and the bearing ring  701  from different materials. Further, the seal ring lug  702  and the bearing ring groove  703  are also on the neutral axis to avoid unwanted twist. As with the ring variations discussed previously, manufacturing tolerances should be adjusted to assure initial contact and sealing between the seal ring  700  and the bore  108 , with the bearing ring sharing increasing load during break-in.