WO2009033115A2 - Segment de piston de haut rendement - Google Patents

Segment de piston de haut rendement Download PDF

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Publication number
WO2009033115A2
WO2009033115A2 PCT/US2008/075516 US2008075516W WO2009033115A2 WO 2009033115 A2 WO2009033115 A2 WO 2009033115A2 US 2008075516 W US2008075516 W US 2008075516W WO 2009033115 A2 WO2009033115 A2 WO 2009033115A2
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WIPO (PCT)
Prior art keywords
ring
pressure
piston ring
piston
gas
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Application number
PCT/US2008/075516
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English (en)
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WO2009033115A3 (fr
Inventor
George Bevan Kirby Meacham
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George Bevan Kirby Meacham
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Application filed by George Bevan Kirby Meacham filed Critical George Bevan Kirby Meacham
Publication of WO2009033115A2 publication Critical patent/WO2009033115A2/fr
Publication of WO2009033115A3 publication Critical patent/WO2009033115A3/fr

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16J—PISTONS; CYLINDERS; SEALINGS
    • F16J9/00—Piston-rings, e.g. non-metallic piston-rings, seats therefor; Ring sealings of similar construction
    • F16J9/12—Details
    • F16J9/20—Rings with special cross-section; Oil-scraping rings

Definitions

  • 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.
  • Reciprocating piston machines including internal combustion engines and gas compressors and are well known and widely used in transportation, industrial, commercial and consumer products.
  • Liquid lubricated compression piston rings are commonly used as sliding seals to reduce gas leakage through the diametral clearance between the cylindrical pistons and the cylindrical bores. Such gas leakage or blow-by reduces the piston effectiveness in all reciprocating piston machines.
  • prior art compression piston rings are effective in reducing blow-by, they are also a source of mechanical friction that consumes power and increases wear. Reduction of ring friction while retaining sealing effectiveness would increase the energy efficiency and service life of a range of equipment essential to the modern economy.
  • 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 the 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.
  • the rings exert a moderate radial elastic force against the bore surfaces to provide a baseline sealing force.
  • 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 grooves between the piston and the rings.
  • This pressurized gas exerts outward radial force on the rings that augments the elastic baseline force and is typically partially balanced by pressure 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.
  • the value of the pressure-driven outward radial force is directly proportional to the axial ring thickness and the differential pressure across the ring.
  • the inward balancing pressure force depends on the details of the contact interface area between the ring and the bore surface. 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.
  • 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. It 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.
  • Reciprocating piston machines including internal combustion engines and gas compressors and are well known and widely used in transportation, industrial, commercial and consumer products.
  • Liquid lubricated compression piston rings are commonly used as sliding seals to reduce gas leakage through the diametral clearance between the cylindrical pistons and the cylindrical bores. Such gas leakage or blow-by reduces the piston effectiveness in all reciprocating piston machines.
  • prior art compression piston rings are effective in reducing blow-by, they are also a source of mechanical friction that consumes power and increases wear. Reduction of ring friction while retaining sealing effectiveness would increase the energy efficiency and service life of a range of equipment essential to the modern economy.
  • 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 the 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.
  • the rings exert a moderate radial elastic force against the bore surfaces to provide a baseline sealing force.
  • 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 grooves between the piston and the rings.
  • This pressurized gas exerts outward radial force on the rings that augments the elastic baseline force and is typically partially balanced by pressure 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.
  • 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 value of the pressure-driven outward radial force is directly proportional to the axial ring thickness and the differential pressure across the ring.
  • the inward balancing pressure force depends on the details of the contact interface area between the ring and the bore surface. 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.
  • 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. It 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.
  • a convex barrel bore contact surface with line contact defines the interface areas exposed to high pressure and low pressure precisely.
  • a tapered 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.
  • L is the axial ring bearing contact width
  • the lubricant film thickness ho 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 ho, and delays the transition to boundary lubrication.
  • 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. It should be noted that the piston velocity is low during boundary layer friction at the stroke ends, minimizing the frictional work. 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.
  • the present invention is directed to designs and methods for reducing friction and wear of conventional and head-land 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.
  • 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.
  • 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.
  • 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.
  • the seal zones and bearing zones are defined by grooves in the outer ring surface that are connected through conduits to gas at the balancing pressure.
  • the conduits may be in the form of holes, slots or grooves. The invention is applicable to both rings that ride in grooves and head-land rings that straddle a land at the piston head.
  • 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.
  • 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.
  • seal zone and bearing zones are defined by relatively deep grooves, rather than subtle features such as taper or ring twist, allows consistent performance over a large wear range.
  • increased area results in a thicker hydrodynamic liquid lubricant film and larger squeeze film lubrication effect that reduces friction and wear at the stroke ends. Altogether, these factors may reduce wear and minimize the need for hard wear-resistant coatings.
  • FIG. 1 illustrates a conventional prior art internal combustion engine piston, cylinder bore and piston ring assembly
  • FIG. 2 illustrates the pressure and force balance on a conventional prior art compression piston ring
  • FIG. 3 illustrates an internal combustion engine piston, cylinder bore and piston ring assembly according to a first embodiment of the present invention
  • FIG. 4 illustrates the pressure and force balance on a compression piston ring according to the first embodiment of the present invention
  • FIG. 5 illustrates a prior art internal combustion engine piston, cylinder bore and piston ring assembly incorporating a head-land compression ring
  • FIG. 6 illustrates the pressure and force balance on a prior art compression headland piston ring
  • FIG. 7 illustrates an internal combustion engine piston, cylinder bore and piston ring assembly incorporating a head-land compression ring according to a second embodiment of the present invention
  • FIG. 8 illustrates the pressure and force balance on a head-land compression ring according to the second embodiment of the present invention
  • FIG. 9 illustrates an internal combustion engine piston, cylinder bore and piston ring assembly incorporating a variation of the first embodiment compression ring of the present invention wherein grooves replace holes as pressure balance passages;
  • FIG. 10 illustrates an internal combustion engine piston, cylinder bore and piston ring assembly incorporating a variation of the second embodiment compression ring of the present invention wherein grooves replace holes as pressure balance passages.
  • Fig. 1 shows a typical prior art piston and ring assembly 100.
  • 101 and second compression ring 102 are disposed in piston grooves 103 and 104 respectively in piston 105 sliding in a bore 106.
  • Oil control ring 107 is disposed in piston groove 108.
  • Compression rings 101 and 102 and oil control ring 107 are manufactured with force-free outside diameters larger than the diameter of the bore 106 and include radial gaps 109. The gaps allow the rings to be expanded elastically for installation into the piston grooves, and to be compressed elastically to fit within the bore. In the compressed condition radial gaps 109 are almost closed to minimize blow-by.
  • Pl is the gas pressure above ring 101
  • P2 is a lower pressure between top compression ring 101 and second compression ring 102.
  • the inside diameter of bore 106 is larger than the outside diameter of piston 105, resulting in diametral clearance 110.
  • the inside diameter 111 of ring 101 is larger than the diameter of the bottom 112 of groove 103, forming an annular volume 113.
  • the lower side 114 of ring 101 is urged against the lower land 115 of groove 103 by the difference between pressures Pl and P2 forming flank seal zone 116 and exposing outer edge 117 of the lower side 113 of ring 101 to lower pressure P2.
  • the upper side 118 of ring 101 is spaced away from upper land 119 of groove 103, forming an annular clearance 120.
  • Annular clearance 120 permits gas at pressure Pl to flow into annular volume 113.
  • the free diameter of ring 101 is larger than the inside diameter of bore 106, causing the ring outer diameter 121 to bear outward against the inner bore surface 122 through its inherent elastic spring action, forming bore seal zone 123.
  • a chamfer 124 is provided between the ring outer diameter 120 and the upper side 118 of ring 101 that defines the width of bore seal zone 123.
  • Liquid lubricant film 125 fills bore seal zone 123 and separates ring outer diameterl21 from inner bore surface 122 under hydrodynamic lubrication conditions.
  • Chamfer 124 on ring 101 provides an annular pressure balance passage 126 between the ring 101 and the inner bore surface 122 that communicates with gas at pressure Pl.
  • FIG. 2 shows prior art compression ring 101 of Fig. 1 under the influence of operating pressures and forces.
  • Gas at pressure Pl fills annular clearance 120, annular volume 113 and annular pressure balance passage 126.
  • Gas at lower pressure P2 contacts the outer edge 117 of the lower side 114 of ring 101.
  • the average static pressures P3 in bore seal zone 123 and P4 in flank seal zone 116 are intermediate between Pl and P2.
  • the net radial force Fr pressing ring 101 against inner bore surface 122 is equal to the vector sum of Fl, the outward radial elastic spring force of ring 101; F2, the outward radial force of high pressure Pl acting on the inside diameter 111 of ring 101; F3, the inward radial force component of high pressure Pl acting on chamfer 124; F4, the inward radial force of intermediate pressure P3 acting on bore seal zone 123; and F8, the radial fiction force between ring 101 and lower flank 115 of groove 103.
  • the net axial force Fz pressing ring 101 against piston lower flank 115 of groove 103 is equal to the vector sum of F5, the, the downward axial force of high pressure Pl acting on the area between the inside and outside diameters of ring 101; F6, the upward axial force of low pressure P2 acting on outer edge 117 of the lower side 114 of ring 101; F7, the upward axial force of intermediate pressure P4 in flank seal zone 116; and Fx, the axial friction force between ring 101 and inner bore surface 122.
  • the magnitude of radial fiction force F8 between ring 101 and lower flank 115 of groove 103 is a function of Fz, and it may be directed either inward or outward depending on the momentary direction of the small radial motions of the ring 101 in the groove 103.
  • Bore seal zone 123 generates an inward force through hydrodynamic or boundary lubrication effects that balances outward radial force Fr.
  • FIG. 3 shows an embodiment of the high efficiency compression ring according to the invention applied to a piston and ring assembly 300. Except for the substitution of the inventive ring 301 for the prior art ring 101, assembly 300 is similar to prior art assembly 100 described with reference to Fig. 1 and Fig. 2. Piston 105 slides in bore 106, and high efficiency compression ring 301 is disposed in piston groove 103. Conventional second compression ring 102 is disposed in piston groove 104, and conventional oil control ring 107 is disposed in piston groove 108. Like second compression ring 102 and oil control ring 107, high efficiency compression ring 301 is manufactured with a force-free outside diameter larger than the diameter of the bore 106 and includes a radial gap 109.
  • the gaps allow the rings to be expanded elastically for installation into the piston grooves, and to be compressed elastically to fit within the bore.
  • Pl is the gas pressure above high efficiency compression ring 301
  • P2 is a lower pressure between ring 301 and second compression ring 102.
  • the inside diameter of bore 106 is larger than the outside diameter of piston 105, resulting in diametral clearance 110.
  • the inside diameter 111 of ring 301 is larger than the diameter of the bottom 112 of groove 103, forming an annular volume 113.
  • the lower side 114 of ring 301 is urged against the lower flank 115 of groove 103 by the difference between pressures Pl and P2 forming flank seal zone 116 and exposing outer edge 117 of the lower side 114 of ring 301 to lower pressure P2.
  • the upper side 118 of ring 301 is spaced away from upper flank 119 of groove 103, forming an annular clearance 120.
  • Annular clearance 120 permits gas at pressure Pl to flow into annular volume 113.
  • the free diameter of ring 301 is larger than the inside diameter of bore 106, causing the ring outer diameter 121 to bear outward against the inner bore surface 122 through its inherent elastic spring action.
  • Ring outer diameter 121 is divided into bore seal zone 123 and a bearing zone 302 by circumferential groove 303.
  • Bearing zone 302 has an upper edge 304 and a lower edge 305.
  • One or more radial holes 306 provide fluid conduits between annular volume 113 and circumferential groove 303.
  • a liquid lubricant film 125 separates bore seal zone 123 and bearing zone 302 from inner bore surface 122 under hydrodynamic lubrication conditions.
  • FIG. 4 shows the high efficiency compression ring 301 described with reference to
  • FIG. 3 under the influence of operating pressures and forces.
  • Gas at pressure Pl fills annular clearance 120, annular volume 113, radial holes 306, and circumferential groove 303.
  • Gas at lower pressure P2 contacts the outer edge 117 of the lower side 114 of ring 301.
  • the average static pressures P3 in bore seal zone 123 and P4 in flank seal zone 116 are intermediate between Pl and P2.
  • the static pressure in bearing zone 302 is Pl, since the both edges of bearing zone 302 are exposed to pressure Pl.
  • the net radial force Fr pressing ring 301 against inner bore surface 122 is equal to the vector sum of Fl, the outward radial elastic spring force of ring 301; F2, the outward radial force of high pressure Pl acting on the inside diameter 111 of ring 301; F3, the inward radial force component of high pressure Pl acting on circumferential groove 303 less the cross section area of the radial holes 306; F4, the inward radial force of intermediate pressure P3 acting on ring bore seal zone 123; F9, the inward radial force of high pressure Pl acting on bearing zone 302: and F8, the radial fiction force between ring 301 and lower flank 115 of groove 103.
  • the net axial force Fz pressing ring 301 against piston lower flank 115 of groove 103 is equal to the vector sum of F5, the, the downward axial force of high pressure Pl acting on the area between the inside and outside diameters of ring 101; F6, the upward axial force of low pressure P2 acting on outer edge 117 of the lower side 114 of ring 301; F7, the upward axial force of intermediate pressure P4 in flank seal zone 116; and Fx, the axial friction force between ring 301 and inner bore surface 122.
  • the magnitude of radial fiction force F8 between ring 301 and lower flank 115 of groove 103 is a function of Fz, and it may be directed either inward or outward depending on the momentary direction of the small radial motions of the ring 301 in the groove 103.
  • Seal zone 123 and bearing zone 302 generate an inward force that balances outward radial force Fr through hydrodynamic or boundary lubrication effects.
  • axial ring friction Fx of high efficiency compression ring 301 is the product of Fr and the instantaneous coefficient of friction.
  • Fr is reduced in both the prior art and inventive rings by reducing the axial width of the bore seal zone 123 to increase the inner radial force by increasing the area between the ring 301 and the inner bore surface 122 acted on by high pressure Pl.
  • reducing the axial width of the bore seal zone 123 in the high efficiency compression ring 301 does not reduce the bearing area.
  • the total effective bearing area in inventive high efficiency compression ring 301 is the sum of the areas of the bore seal zone 123 and the bearing zone 302.
  • the radial force Fl depends only on the area of bore seal zone 123, and not the area of bearing zone 302. The result is that, unlike in prior art ring 101, radial force Fl and the total bearing area of ring 301 may be adjusted independently. This allows robust rings with large bearing area that combine low friction and low wear.
  • the width of bore seal zone 123 is set to generate the minimum radial force Fr needed to overcome the groove friction force F8 and to assure that the ring will dynamically follow and contact the bore throughout the piston motion to minimize blow-by. Bore seal zone 123 may be narrow without incurring excessive wear, since bearing zone 302 is made relatively wide and carries much of radial force Fr. As discussed above, the hydrodynamic coefficient of friction is independent of the width of seal zone 123 and bearing zone 302.
  • Increased bearing width does, however, increase the thickness of hydrodynamic liquid lubricant film 125. This is beneficial, since it delays the transition from hydrodynamic to boundary layer lubrication at the stroke end, decreasing friction and wear. Increased width also increases the squeeze film lubrication effect that can further delay or even prevent a transition to boundary lubrication at the stroke end. Finally, increased width reduces radial ring wear during periods of boundary lubrication at the stroke end by reducing the contact pressure and distributing the wear over a larger contact area. 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.
  • inventive ring 301 is a robust design that reduces the need to compromise between friction, blow-by and wear to meet the demands of a particular application.
  • FIG. 5 shows a typical piston and ring assembly 500 incorporating a prior art headland compression ring 501.
  • Such rings are designed to reduce crevice volume and the associated quench zones typical of compression rings disposed in grooves that reduce combustion efficiency and increase hydrocarbon emissions.
  • Top flange 502 and bottom flange 503 of head-land ring 501 extend radially inward from cylindrical body 504 and straddle the piston land 505 of piston 506.
  • Flange 507 extends radially outward from cylindrical body 504 to form the outside diameter 508 of head-land ring 501.
  • inward top flange 502 and outward flange 507 are approximately equal, and inward top flange 502 and outward flange 507 are approximately coplanar.
  • the total top pressure area 509 of head-land ring 501 is the annular area between the ring outside diameter 508 and the inside diameter 510 of inward top flange 502 projected on a plane perpendicular to the axis of bore 106.
  • Second compression ring 102 and oil control ring 107 are disposed in piston grooves 104 and 108 respectively in piston 506 sliding in bore 106.
  • Head-land ring 501, second compression rings 102 and oil control ring 108 are manufactured with force-free outside diameters larger than the diameter of the bore 106 and include radial gaps 109.
  • the gaps allow the rings to be expanded elastically for installation into the piston, and to be compressed elastically to fit within the bore.
  • Pl is the gas pressure above ring 501
  • P2 is a lower pressure between head-land ring 501 and second compression ring 102.
  • the inside diameter of bore 106 is larger than the outside diameter of piston 506, resulting in diametral clearance 110.
  • the inside diameter 513 of ring 501 between flanges 502 and 503 is larger than the outer diameter 514 of piston land 505, forming an annular volume 515.
  • the upper inside surface 516 of ring 501 is urged against the upper face 517 of piston land 505 by the difference between pressures Pl and P2, forming flank seal zone 518.
  • the lower inside surface 519 of ring 501 is spaced away from the lower face 520 of piston land 505, forming an annular clearance 521.
  • Annular clearance 521 permits gas at pressure P2 to flow into annular volume 515.
  • Pressure P2 acts on the annular area 522, where the inner diameter of this area is outer diameter 514 of piston land 505 and the outer diameter of this area is the outside diameter 508 of ring 501, both projected on a plane perpendicular to the axis of bore 106.
  • the free diameter of ring 501 is larger than the inside diameter of bore 106, causing the head-land ring outer diameter 508 to bear outward against the inner bore surface 122 through its inherent elastic spring action to form seal zone 123.
  • a chamfer 124 is optionally provided between the ring outer diameter 508 and the upper side 523 of ring 501 that reduces the width of bore seal zone 123.
  • Chamfer 124 on ring 501 provides the annular pressure balance passage 126 between the ring 501 and the inner bore surface 122 that communicates with gas at pressure Pl.
  • Liquid lubricant film 125 separates bore seal zone 123 from inner bore surface 122 under hydrodynamic lubrication conditions.
  • FIG. 6 shows prior art head-land compression ring 501 of Fig. 5 under the influence of operating pressures and forces.
  • Gas at higher pressure Pl acts on total top area 509, inside diameter 510 of inward top flange 502, and annular pressure balance passage 126.
  • Gas at lower pressure P2 fills annular volume 511, annular clearance 521, and annular volume 515, with the net effect that P2 acts on annular area 522 of head- land ring 501.
  • the average static pressures P3 in bore seal zone 123 and P4 in flank seal zone 517 are intermediate between Pl and P2.
  • the net radial force Fr pressing ring 501 against bore 106 is equal to the vector sum of Fl, the outward radial elastic spring force of ring 101; F2, the outward radial force of high pressure Pl acting on the inside diameter 510 of inward top flange 502; F3, the inward radial force component of high pressure Pl acting on chamfer 124; F4, the inward radial force of intermediate pressure P3 acting on bore seal zone 123; and F8, the radial fiction force between inside upper face 516 of ring 501 and upper surface 517 of piston land 505.
  • the net axial force Fz pressing ring 501 against upper face 517 of piston land 505 is equal to the vector sum of F5, the downward axial force of high pressure Pl acting on total top area 509 of ring 501; F6, the upward axial force of low pressure P2 acting on annular area 522 of ring 501; F7, the upward axial force of intermediate pressure P4 in P4 in flank seal zone 518; and Fx, the axial friction force between ring 501 and inner bore surface 122.
  • the magnitude of radial fiction force F8 between ring 501 and upper face 517 of piston land 505 is a function of Fz, and it may be directed either inward or outward depending on the momentary direction of the small radial motions of the ring 501 on the piston land 105.
  • Seal zone 123 generates an inward force through hydrodynamic or boundary lubrication effects that balances outward radial force Fr.
  • Prior art head-land ring 501 does not allow gas at high pressure Pl into crevices in which the flame is quenched, avoiding reduced combustion efficiency and increased hydrocarbon emissions.
  • the seal zone 123 is at the piston head, eliminating the annular crevice between a conventional piston and the bore above the compression ring.
  • the upper inside surface 516 of ring 501 is urged against the upper face 516 of piston land 505, forming flank seal zone 518 that effectively prevents gas from entering crevices between piston 506 and ring 501. Instead, the crevices between piston 506 and ring 501 are filled with gas at low pressure P2 that is not part of the combustion process.
  • FIG. 7 shows an embodiment of the high efficiency head-land compression ring according to the invention applied to a piston and ring assembly 700. Except for the substitution of the inventive high efficiency head-land ring 701 for the prior art headland ring 501, assembly 700 is similar to prior art assembly 500 described with reference to Fig. 5 and Fig. 6, and includes the advantages of reduced crevice volume and quench zones.
  • Top flange 502 and bottom flange 503 of high efficiency head- land ring 701 extend radially inward from cylindrical body 504 and straddle the piston land 505 of piston 506. The outside diameter of cylindrical body 504 forms the outside diameter 508 of high efficiency head-land ring 701.
  • the total top pressure area 509 of the head-land ring top surface 702 is the annular area between the ring outside diameter 508 and the inside diameter 510 of inward top flange 502, both projected on a plane perpendicular to the axis of bore 106.
  • Second compression ring 102 and oil control ring 107 are disposed in piston grooves 104 and 108 respectively in piston 506 sliding in bore 106.
  • High efficiency head-land ring 701, second compression rings 102 and oil control ring 108 are manufactured with force-free outside diameters larger than the diameter of the bore 106 and include radial gaps 109. The gaps allow the rings to be expanded elastically for installation into the piston, and to be compressed elastically to fit within the bore.
  • Pl is the gas pressure above ring 701, and P2 is a lower pressure between head- land ring 701 and second compression ring 102.
  • the inside diameter of bore 106 is larger than the outside diameter of piston 506, resulting in diametral clearance 110.
  • the inside diameter 513 of ring 701 between flanges 502 and 503 is larger than the outer diameter 514 of piston land 505, forming an annular volume 515.
  • the upper inside surface 516 of ring 701 is urged against the upper face 517 of piston land 505 by the difference between pressures Pl and P2, forming flank seal zone 518.
  • the lower inside surface 519 of ring 701 is spaced away from the lower face 520 of piston land 505, forming an annular clearance 521.
  • Annular clearance 521 permits gas at pressure P2 to flow into annular volume 515.
  • Pressure P2 acts on the annular area 522, where the inner diameter of this area is outer diameter 514 of piston land 505 and the outer diameter of this area is the outside diameter 508 of piston ring 701, both projected on a plane perpendicular to the axis of bore 106.
  • the free diameter of ring 701 is larger than the inside diameter of bore 106, causing the head-land ring outer diameter 510 to bear outward against the inner bore surface 122 through its inherent elastic spring action.
  • Ring outer diameter 508 is divided into bore seal zone 123 and a bearing zone 302 by circumferential groove 303.
  • Bearing zone 302 has upper edge 304 and lower edge 305.
  • One or more radial holes 306 provide fluid conduits between annular volume 515 and circumferential groove 303.
  • the upper edge 703 of groove 303 is approximately in the plane of upper inside surface 516 of ring 701.
  • a chamfer 124 is optionally provided between the ring outer diameter 121 and the top surface 702 of ring 701 that reduces the width of bore seal zone 123.
  • Chamfer 124 on ring 701 provides an annular pressure balance passage 126 between the ring 701 and the inner bore surface 122 that communicates with gas at pressure Pl.
  • Liquid lubricant film 125 separates bore seal zone 123 and bearing zone 302 from inner bore surface 122 under hydrodynamic lubrication conditions.
  • FIG. 8 shows prior art high efficiency head-land compression ring 701 of Fig. 7 under the influence of operating pressures and forces.
  • Gas at higher pressure Pl acts on total top area 509, inside diameter 510 of inward top flange 502, and annular pressure balance passage 126.
  • Gas is at lower pressure P2 in annular volume 515, annular clearance 521, radial holes 306, circumferential groove 303, and bearing zone 302, with the net effect that P2 acts on annular area 522 of head-land ring 501.
  • the average static pressures P3 in bore seal zone 123 and P4 in flank seal zone 517 are intermediate between Pl and P2.
  • the net radial force Fr pressing ring 701 against inner bore surface 122 is equal to the vector sum of Fl, the outward radial elastic spring force of ring701; F2, the outward radial force of high pressure Pl acting on the inside diameter 510 of inward top flange 502; F3, the inward radial force component of high pressure Pl acting on chamfer 124; F4, the inward radial force of intermediate pressure P3 acting on bore seal zone 123; and F8, the radial fiction force between inside upper face 516 of ring 701 and upper surface 517 of piston land 505.
  • the net axial force Fz pressing ring 701 against upper face 516 of piston land 505 is equal to the vector sum of F5, the downward axial force of high pressure Pl acting on total top area 509 of ring 501; F6, the upward axial force of low pressure P2 acting on annular area 521 of ring 501; F7, the upward axial force of intermediate pressure P4 in P4 in flank seal zone 517; and Fx, the axial friction force between ring 701 and inner bore surface 122.
  • the magnitude of radial fiction force F8 between ring 701 and upper face 517 of piston land 505 is a function of Fz, and it may be directed either inward or outward depending on the momentary direction of the small radial motions of the ring 701 on the piston land 105.
  • Seal zone 121 and bearing zone 302 generate a net inward force through hydrodynamic or boundary lubrication effects that balances outward radial force Fr.
  • high efficiency head-land ring 701 does not allow gas at high pressure Pl into crevices in which the flame is quenched, thereby avoiding reduced combustion efficiency and increased hydrocarbon emissions.
  • the wear, friction and lubrication characteristics of high efficiency head-land compression ring 701 are similar to those of high efficiency compression ring 301 described with respect to Fig. 3 and Fig. 4.
  • Axial ring friction Fx is reduced at all parts of the stroke by controlling and minimizing radial force Fr, while simultaneously increasing the ring bearing area to extend the hydrodynamic lubrication into the stroke ends, thereby reducing the friction and wear associated with boundary lubrication.
  • the static pressure in bearing zone 302 of high efficiency head-land compression ring 701 is the low pressure P2 rather that the high pressure Pl, but this does not affect the hydrodynamic force generated by the bearing.
  • radial force Fl and the total bearing area may be adjusted independently. This allows robust head-land rings with large bearing area that combine low friction and low wear, and reduces the need to compromise between friction, blow-by and wear to meet the demands of a particular application.
  • FIG. 9 shows a variation 900 of high efficiency compression ring 301 described with respect to Fig. 3 and Fig. 4.
  • Radial holes 306 that form fluid conduits between gas at high pressure Pl in annular volume 111 to pressurize circumferential groove 303 are eliminated, and one or more axial flow grooves 901 are formed in the surface of bearing zone 302. These grooves provide fluid conduits between upper edge 304 and lower edge 305 of bearing zone 302 so that gas at high pressure Pl flows through axial grooves 901, and pressurizes circumferential groove 303.
  • Fig. 10 shows a variation 1000 of high efficiency head-land compression ring 701 described with respect to Fig. 7 and Fig. 8.
  • Radial holes 304 that form fluid conduits between gas at lower pressure P2 in annular volume 114 to pressurize circumferential groove 303 are eliminated, and one or more axial flow grooves 901 are formed in the surface of bearing zone 302. These grooves provide fluid conduits between upper edge 304 and lower edge 305 of bearing zone 302 so that gas at lower pressure P2 flows through axial grooves 900, and sets the pressure of circumferential groove 303.
  • the axial grooves 901 serving as flow conduits in variations 900 and 1000 shown in Figures 9 and 10 may prove more resistant to clogging by particulates, e.g. carbon or metallic wear products, than drilled radial holes 304. They also facilitate alternative production process, e.g. embossing or broaching, that may be more economical than the drilling process required for radial holes.
  • piston rings made of a variety of known metal, ceramic, composite and polymer solid materials utilizing known manufacturing processes including but not limited to casting, forging, molding, conventional machining, electrical discharge machining, laser machining and sintering. Further, coating materials and techniques known to be useful for piston rings may be employed. Similarly, it may incorporate known piston ring design features such as ring twist, taper, chamfers, and fillets.
  • the rings may be used in combination with pistons incorporating known features including groove coatings, groove taper, groove tilt, and pressure equalization passages.
  • the embodiments were chosen and described in order to best explain the principle of the invention and its practical applications to thereby enable others skilled in the art to best utilize the invention in its various embodiment and with various modifications as are suited to the particular use contemplated. It is intended that the invention be defined by the following claims.
  • "conventional-type” ring refers to a piston ring of generally rectangular cross section disposed in a piston groove
  • head-land-type refers to a piston ring of generally U-shaped cross section disposed over a piston land.

Abstract

La présente invention concerne un segment de piston dans lequel un gaz de segment de piston lubrifié liquide amélioré réalise une obturation avec un frottement et une usure réduits en comparaison des segments de la technique antérieure, à utiliser dans des moteurs à combustion interne, des pompes à gaz et des compresseurs à gaz. Ces segments améliorés permettent un ajustement indépendant des paramètres régulant le frottement et l'usure, et éliminent le compromis à réaliser entre faible frottement et faible usure typique des segments de la technique antérieure.
PCT/US2008/075516 2007-09-08 2008-09-06 Segment de piston de haut rendement WO2009033115A2 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
US96786007P 2007-09-08 2007-09-08
US60/967,860 2007-09-08

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WO2009033115A3 WO2009033115A3 (fr) 2009-06-25

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Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102012002443A1 (de) * 2012-02-08 2013-08-08 Mahle International Gmbh Kolbenring für einen Verbrennungsmotor

Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3917290A (en) * 1972-11-28 1975-11-04 Robert Geffroy Assembly comprising a piston groove, a piston ring and an inertia ring sliding in a cylinder
US4128250A (en) * 1974-12-21 1978-12-05 Mahle Gmbh Pistons and piston rings
JP2003097712A (ja) * 2001-09-27 2003-04-03 Mitsubishi Heavy Ind Ltd ピストンリング及びピストン、並びにピストン機関
JP2006153198A (ja) * 2004-11-30 2006-06-15 Toyota Motor Corp オイルリング

Patent Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3917290A (en) * 1972-11-28 1975-11-04 Robert Geffroy Assembly comprising a piston groove, a piston ring and an inertia ring sliding in a cylinder
US4128250A (en) * 1974-12-21 1978-12-05 Mahle Gmbh Pistons and piston rings
JP2003097712A (ja) * 2001-09-27 2003-04-03 Mitsubishi Heavy Ind Ltd ピストンリング及びピストン、並びにピストン機関
JP2006153198A (ja) * 2004-11-30 2006-06-15 Toyota Motor Corp オイルリング

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102012002443A1 (de) * 2012-02-08 2013-08-08 Mahle International Gmbh Kolbenring für einen Verbrennungsmotor
US9279499B2 (en) 2012-02-08 2016-03-08 Mahle International Gmbh Piston ring for an internal combustion engine

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