have better high temperature wear resistance with a sacrifice in low temperature flexibility. PTFE, a thermoplastic rather than an elastomer, has a wide temperature range and is resistant to almost all fluids. It is difficult to process and is usually employed as assembled seals. Butyl, epichlorhydrin, an ethylene-propylene terpolymer (EPDM) are used in special purpose seals. Packing Seals Mechanical shaft packings include compression packing, automatic or lip packing, and squeeze packing. Compression packings are a pliable material compressed between the throat and gland of a stuffing box for reciprocating, oscillating, and rotating applications. Leakage in dynamic applications is usually on the order of 50 to 500 mᐉ/hr, but may be essentially zero in semistatic valve stem applications. Automatic packings utilize a flexible lip energized by the contained fluid pressure. Employed primarily for reciprocating applications, heat dissipation problems restrict rare rotating applications to speeds below 1 m/sec (200 ft/min). Squeeze packings utilize precision-molded elastomer rings, such as the O-ring, installed in precisely machined grooves (glands) on cylinders, pistons, or rods in hydraulic or pneu- matic devices. 25,26 Squeeze packings are most frequently used in reciprocating service or in low-speed oscillating applications such as valve stems. Rotary applications are recommended only under well-lubricated low speed conditions, 1.75 to 4 m/sec (350 to 800 ft/min). None of these packing devices are bearings. Side loads due to out-of-round parts, warped shafts, or poor bearing supports will cause rapid wear and inadequate sealing. Compression Packing The soft packing, jamb packing, or compression packing, Figure 18, is the most common fluid seal. It consists of a number of deformable packing rings or a long rope-like material spiral wrapped around the shaft or rod, compressed by the gland to seal against the housing bore and shaft. Leakage on the order of 0.01 mᐉ/hr/m/kPa (0.0018 mᐉ/hr-in psi) is necessary to lubricate and cool the packing. Leakage from a compression packing will be approximately 5 to 100 times that from a mechanical face seal under the same service conditions and friction loss will be about three times greater. Compression packing has the advantage of being replaceable without disassembly of equipment and a gradual leakage increase usually Volume II605 FIGURE 18. Typical pomp stuffing box with compression packing. [1] Shaft finish = 0.25 to 0.50 µm (10 to 20 µin.) CLA; shaft hardness = Rockwell C-50; shaft runout should not exceed 0.025 mm (0.001 in.) TIR. [2] Bore finish = 1 to 1.5 µm (40 to 60 µin.) CLA. [3] Rings nearest gland are deformed most; approximately 70% of wear under first 30% of packing. [4] Harder end rings are sometimes used at gland and at throat. [5] Packing length ~ – 1.5 D. [6] Packing radial thickness ~ _ 0.15 to 0.3 D. [7] Throat clearance 0,2 to 0.4 mm (0.008 to 0.015 in); 0.8 mm maximum. [8] Gland-to-bore clearance 0.125 to 0.25 mm (0.005 to 0.010 in.). [9] Gland-to-shaft clearance 0.4 to 0.8 mm (0.015 to 0.030 in.). [10] Tap locations for lantern gland inlet. [11] Lantern ring. 581-622 4/10/06 6:03 PM Page 605 Copyright © 1983 CRC Press LLC provides adequate warning of impending failure. While initial cost of compression packings is lower, their periodic maintenance and adjustment for wear and loss of packing volume frequently swing total cost in favor of mechanical seals. Compression packings are used extensively in rotary applications such as pumps up to about 15 m/sec for pressures up to 1000 kPa (145 psi) and valve stems under semistatic conditions up to 34,500 kPa (5000 psi). Compression packing are sometimes used for sealing reciprocating shafts but they have the disadvantage of high friction. Figure 19 shows representative designs and the most frequently used materials. Repre- sentative packings, lubricants, temperature limits, and applications are shown in Table 12. Soft packing, usually square cross section rings or long continuous pieces which can be 606 CRC Handbook of Lubrication FIGURE 19. Typical soft packing and commonly used materials: (a) spiral-wrapped metal foil over reinforced braided asbestos core; (b) crumpled metal foil, graphited; (c) cotton duck laminated with synthetic rubber; (d) lead wire reinforced flax braid over synthetic rubber core; (e) folded and wrapped asbestos fabric, soft rubber core at housing bore; and (f) graphite foil wound around shaft and then compressed. COMMONLY USED MATERIALS 581-622 4/10/06 6:03 PM Page 606 Copyright © 1983 CRC Press LLC For compression packings, it is best to use die-formed rings which may be purchased as a set or prefabricated by the user in a mold of correct dimensions. These rings minimize gland take-up during break-in, enhance extrusion resistance, reduce the break-in period, tend to exclude abrasives, and allow sealing at higher pressures. The ring OD may be slightly oversize to provide good housing bore fit. Atypical packing set may use very dense “anti- extrusion” rings at the throat and gland with intermediate rings graded from soft near the throat to hard near the gland. 27 Alantern ring, Figure 18, is frequently used in compression packings for rotary appli- cations, especially at high pressures and temperatures. The lantern ring has an H cross section and is made of rigid material such as brass, aluminum, stainless steel, or PTFE. The ring is adjacent to openings in the stuffing box wall for injecting coolants or lubricants, and a discharge can be provided on the opposite side of the housing. The lantern ring can also be used to (1) introduce fluid from pump discharge when pump suction is subatmospheric to prevent air leaking in, and (2) introduce a clean external buffer liquid to seal against abrasives, slurries, toxic liquids, and gases. The buffer fluid pressure should be about 20 to 70 kPa (3 to 10 psi) above the pump suction. The lantern ring is usually located about midway in the packing set but its exact location may be dictated by suction pressure, lubricant viscosity, or buffer fluid pressure. Automatic Packing Pressure-energized lip-type automatic packings, the most widely used seal in the high pressure hydraulic and pneumatic field, are generally installed with a very small interference. Contact force and area increase with fluid pressure, improving the seal. Used almost ex- clusively for reciprocating applications, contact force, area, and friction on an unpressurized return stroke are lower than on the pressure stroke and produce a “breathing” action that helps lubricate the seals. The friction of automatic packing is approximately proportional to pressure up to about 7000 kPa (1015 psi). Above this, the rate of friction increase with pressure decreases and becomes quite small at about 14,000 kPa (2030 psi). 28 Automatic packings are depicted in Figure 20 in order of increasing pressure limits. They are available in a wide variety of homogeneous elastomers or fabric-reinforced compositions. Cup and flange packing — These are the simplest designs, require a minimum of space, and are easily installed (Figure 21). The flange packing OD and cup packing ID are sealed by mechanical compression, which limits maximum operating pressure to approximately 3500 kPa (500 psi). Excessive tightening of the inside follower tends to crush and extrude the cup packing against the cylinder wall, which causes high friction, wear, and reduced sealing effectiveness. Similar crushing of the flange packing may result from gland over- tightening. Cup and flange packings are less effective seals than U- or V-rings but are frequently used because of space limitations. Leather continues to be much used for flange packing along with various synthetic rubbers, PTFE, nylon, and other plastics. Fabric- reinforced elastomers greatly reduce problems with mechanical clamping. U-ring packing — These low-friction packings of leather, elastomer, or fabric-reinforced elastomer are used singly in continuous (nonsplit) rings. They are infrequently used in tandem. U-rings are chiefly employed as piston seals but can be arranged in glands. In double-acting piston seals, the U-ring must be used heel-to-heel. A lip-to-lip arrangement will create a pressure trap and cause rapid seal wear and failure. Homogeneous U-rings in Shore A hardness of 70 can be used up to about 10,000 kPa (1450 psi) in precision machined parts. Maximum radial clearance should be about 0.075 mm (0.003 in.). For higher pressures or for applications with excess clearance, harder U-rings up to Shore A of 90 and/or fabric- reinforced rings should be used. U-rings with metal-reinforced bases have been used up to 35,000 kPa (5100 psi). Some proprietary U-ring designs having long thick-walled static sealing lips can be installed with enough interference to make pedestal rings unnecessary. 608 CRC Handbook of Lubrication 581-622 4/10/06 6:03 PM Page 608 Copyright © 1983 CRC Press LLC cut, the joints spaced at 120°, to simplify replacement without machine disassembly. V- rings are available in leather, homogeneous elastomers, fabric-reinforced elastomers, and PTFE. Split rings are usually fabric reinforced. Homogeneous rings are used up to about 20,000 kPa (2900 psi). At pressures around 35,000 kPa (5100 psi), homogeneous rings can be mixed with leather or PTFE rings, or a combination of different hardness rings can be used with softer, more leak-tight rings placed nearest the high pressure. At pressures above 45,000 kPa (6500 psi), endless fabric-reinforced elastomer or PTFE rings are common, and thin metal separators frequently support each pressure ring. V-rings can be used as piston seals but are more commonly used in rod seal glands. V-rings can be designed to withstand almost 45,000 kPa (6500 psi) per ring, but this practice results in poor seal life. Three rings are usually the fewest employed even at modest pressures. At 35,000 kPa (5100 psi), a typical packing set would have five or six rings. The male and female support rings are usually made from the same material as the pressure rings when used at low pressures, less than 20,000 kPa (2900 psi). For higher pressures, support rings are available in PTFE, rockhard duck and rubber, metal and phenolic. Installation — Industry standardization is greater for automatic packing than for any other seal type. Many failures result from a disregard of design and dimensional information provided by the packing manufacturer. Aproblem common to lip-type automatic packings is extrusion due to high pressure and excess clearance. Metal surfaces in sliding contact with automatic packing should be finished to 0.2 to 0.4 µm (8 to 16 µin.). Finish should not be smoother than about 0.13 µm (5 µin.) because slight roughness helps retain lubricant. The static surface in contact with the packing should be finished to 0.8 µm (32 µin.). Squeeze Packing Squeeze packings are made in several shapes, in a large number of standardized sizes, 25 and from over a dozen elastomers with hardness ranging from 10 to 100 Shore A. 21 These seals, Figure 23, are low in cost, require minimum space, are easy to install, require no adjustment, seal in both directions, have low friction, can be used as piston or gland seals, can be selected for compatibility with a wide range of fluids, and are readily available for industrial, aerospace, and military applications. Squeeze rings, though simple in form, are made with closely held diametral and cross section tolerances. To ensure long life and effective sealing, recommended groove dimensions, surface finishes, and diametral clear- ances must be carefully followed. 610CRC Handbook of Lubrication FIGURE 22. V-ring automatic gland seal. 581-622 4/10/06 6:03 PM Page 610 Copyright © 1983 CRC Press LLC diameter slightly smaller than the O-ring OD and the groove diameter is slightly smaller than the O-ring ID. With changes in pressure and direction, a momentary leak occurs as the ring moves from one side of the groove to the other. Since this design is primarily for low-pressure pneumatic service, about 1380 kPa (200 psi), this slight leakage is generally acceptable. This arrangement can also be used in low-pressure liquid service if a few drops of leakage per cycle can be tolerated. Dynamic O-ring seals are used primarily for well-lubricated reciprocating service. With proper design, however, they can be employed in low-speed rotary service at pressures up to about 5500 kPa (800 psi). The gland for rotary applications compresses the O-ring about 5% circumferentially. Its depth is only slightly less than the O-ring cross-section, so there is little radial squeeze. Rotary seals are not put in tension around the shaft because most elastomers if heated by friction while under tensile stress will contract. This contraction, the Gow-Joule effect, causes further contact load, increased friction and temperature, and rapid failure. O-rings and other squeeze packings are made from a large number of elastomers in hardnesses from about 55 to 90 Shore A. Astandard O-ring with a hardness of 60 will seal pressures in dynamic applications to about 1750 kPa (250 psi) and about 10,500 kPa (1500 psi) with a 90 hardness. Higher pressures, up to about 20,700 kPa (3000 psi), require backup rings to prevent ring extrusion. T-ring shape can be used up to about 138,000 kPa (20,000 psi). Table 13 gives some characteristics of the most widely used elastomers. CONTROLLED CLEARANCE SEALS Hydrodynamic Seals While mechanical face seals often function with separation of the sealing surfaces because of static or dynamic pressure forces, 30 controlled close clearance seals provides a definite sealing surface separation during normal operation. The hydrodynamic seal shown in Figure 26 was designed for gas, but hydrodynamic seals can also be used for liquids. Essentially, thesealing ring interface is an ordinary mechanical face seal with a fluid film bearing geometry added to give positive separation of the surfaces. The self-acting lift pads have pockets about 10 to 25 µm (0.0005 to 0.001 mᐉ) deep and pocket-to-land width ratios in thecircumferential direction of about 2:1. Axial and radial grooves keep pressure the same around each pad. During seat rotation, high-pressure gas is dragged into the pad and com- pressed as it passes over the step at the end of the pad. This creates lift forces that separate the primary seal ring and rotating seat. 612CRC Handbook of Lubrication FIGURE 25. O-ring dynamic seal gland detail. Surface finishes: X = 0.254 to 0.508 µm (10 to 20 µin.) CLA; NOTE: do not use less than 0.127 µm (5 µin.); Y = 0.8 µm (32 µin.) CLA; Z = 0.8 µm (32 µin.) CLA without backup rings, 1.6 µm when used with backup; and B = groove shown for no backup ring. If ring is employed use supplier’s recommendation for B. 581-622 4/10/06 6:03 PM Page 612 Copyright © 1983 CRC Press LLC The pressure drop and leakage occur across the sealing dam of the sealing ring. The fluid film bearing also contributes high film stiffness such that the seal ring can dynamically track seal seat motion. This is especially important in high-speed applications where runout could not otherwise be tolertated Aspiral groove pattern can be applied on the seal face to operate in a manner similar to the lift pads. 31 with a wide radial face, pumping action of the spiral grooves can result in zero net leakage under ideal conditions. Hydrosatic Seals There are two kinds of hydrosatic close clearance seals: self activated and externally pressurized. Figure 27 shows a self-activated hydrosatic seal with a shallow radial step approximately at midface. In case A(normal design separaion), the hydrostatic seprating is in equilibrium ith the seal closing ( hydrostatic pressure) force as shown. If face separation decreaes or increases, a restoring force develops due to the change in pressure profile as shown in B and C. Similar performance and stability can be achived with a gradually converging face sepration and high leakage. Alternatively, a midface pocket in a flat-faced seal can be connected to the high-pressure side through an additional channel offering resistance to flow. With approprite geometries, pressure profiles are similar to those in Figure 27. Instability problems sometimes occur with gases when operating with relatively large face sepration and high leakage. Generally, these seals are used in high pressure differential applictions. Rotation usually has a negligible effect in these cases (rotational speed is too low and separation is too high for significant hydodynamic effects). An externally pressurized hydrostatic seal is shown in figure 28. Under all conditions of opertion, the buffer pressure must be higher than the sealed pressure. The buffer fluid overpressure may be relatively low, 15 to 35 kPa (2.5 to 5 psi), and is usually dicated by the control system employed. Where abrasives are present in the sealed fluid, the buffer fluid flushes abrasives away from the sealing interface. This principle is also used for sealing toxic fluids. if the buffer fluid is not compatible with the sealed fluid, a more complex seal system is required. Hydrodynamic and hydrostatic concepts are combined in a hybrid seal in figure 29. At zero and low pressures, hydrodynamic pumping allows operating without face cotact. Although the seal gap does incease with speed, the increase is moderate throuhghout a large 614CRC Handbook of Lubrication FIGURE 27. Self-activated hydrostatic face seal. A = seal opening pressure distribution at equilibrium h, B at small h, C at large h. 581-622 4/10/06 6:03 PM Page 614 Copyright © 1983 CRC Press LLC ments and still behave as a close clearance seal. Multiple short rings can be staged for better sealing and to accommodate shaft misalignment. In high-temperature applications, thermal expansion of the bushing must match that of the shaft. The basic mass flow equations for incompressible constant area parallel flow 34 are Laminar (11) Turbulent (12) The flow model for a bushing seal is shown in Figure 31. Since flow path width is W = 2πR, laminar concentric annular flow between the cylindrical surfaces is (13) For an eccentric annular film, film thickness h = h m (1 + ⑀ cos θ ), where θ is reckoned from the position at which h = h minimum , and ⑀ = e/h m , Equation 13 for laminar flow becomes: (14) When the annuius is fully eccentric, ⑀ = 1 and the factor (1 + 1.5 ⑀ 2 ) becomes 2.5. Substituting 2πR for W in Equation 12 for turbulent concentric flow: (15) The fully eccentric correction factor for full turbulence is 1.315, where M · = mass velocity, L = length of flow path, W = width of flow path, h = film thickness, h m = mean film thickness, P = pressure, R = radius, e = eccentricity, µ = absolute viscosity, and ρ = fluid density. 616 CRC Handbook of Lubrication FIGURE 31. Flow model for bushing seal. 581-622 4/10/06 6:03 PM Page 616 Copyright © 1983 CRC Press LLC FIXED-GEOMETRYCLEARANCE SEALS Buffered Bushing Seal Bushing seals depend on small clearances between relatively moving surfaces and are commonly used to limit leakage of liquids. They are frequently used as shown in Figure 32 with process fluid leakage being prevented by a reverse leak of buffer fluid. To minimize ingress of buffer fluid, the primary bushing pressure differential, (p b – P p ), should be small. On the other hand, a process gas may leak against a small primary bushing pressure gradient. While the buffered seal arrangement generally requires an extensive system of piping, pumps, heat exchangers, separators, and controls, the seal has much potential for large systems, particularly those containing hazardous fluids. Labyrinth Seal Labyrinth seals, which comprise a series of flow restrictions as shown in Figure 33, capitalize on entrance and exit losses and turbulence to minimize leakage flow. Their ef- fectiveness is highly dependent on the annular clearance between the rotating shaft and stationary housing. The labyrinth seal has a long history and is widely used to minimize steam or gas leakage when direct contact and wear between sealing members is not feasible. Leakage rates are relatively high compared to other seal types. Analysis of the labyrinth seal has generally considered the labyrinth as an orifice, 35 or as turbulent pipe flow. The actual process lies somewhere between. Using the former approach, Egli 36 derived the leakage equation and curves in Figure 34, where A = leakage area, α = contraction factor, φ = flow function, γ = carryover factor, M · = mass velocity, ρ 1 = entrance fluid density, and p 1 = entrance fluid pressure. SEALS USING SPECIALIZED CONTROLOF FLUID Freeze Seal Freeze seals have been used primarily by the nuclear industry as stem seals for valves handling liquid sodium, potassium, and lead (Figure 35). Basically, liquid metal solidifies in the annulus around the shaft and acts as the seal. In operation, frictional or other heat causes a thin fluid film to develop between mating parts. Properly designed, the freeze seal will have a starting torque no greater than a typical packing seal and lower running power. Atypical gap is 0.76 mm (30 mil): small enough to prevent extrusion of a solid sodium Volume II617 FIGURE 32. Simple buffered bushing seal. (From Stair, W. K., Liquid buffered bushing seals for large gas circulators, Paper C5, presented at 1st Int. Conf. Fluid Sealing, BHRA, Fluid Engineering, Cranfield, Bedford, England, April 1961.) 581-622 4/10/06 6:03 PM Page 617 Copyright © 1983 CRC Press LLC REFERENCES 1. Bernd, L. H., Survey of the theory of mechanical seals. I. Characteristics of seals, Lubr. Eng., 24(10), 479, 1968. 2. API, Centrifugal Pumps for General Refining Services, API Standard 610, 5th ed., American Petroleum Institute, Washington, D.C., March 1971. 3. Ludwig, L. P. and Greiner, H. F., Designing mechanical face seals for improved performance. I. Basic configurations, Mech. Eng., 100(11), 38, 1978. 4. Anon., Guide to Modern Mechanical Sealing, 6th ed., Durametallic. Corporation, Kalamazoo, Mich., 1971. 5. Austin, R. M., Nau, B. S., Guy, N., and Reddy, D., The Seal Users Handbook, 2nd ed., BHRA Fluid Engineering, Cranfield, Bedford, England, 1979. 6. Stevens, J. B., Pace seals — metal bellows types, Mach. Design, 41(14), 32, 1969. 7. Stair, W. K. and Ludwig, L. P., Energy conservation through sealing technology, Lubr. Eng., 34(11), 618, 1978. 8. Schoenherr, K., Materials in End-Face Mechanical Seals, No. 63-WA-254, American Society of Me- chanical Engineers, New York, 1963, preprint. 9. Lymer, A. and Greenshield, A. L., Thermal aspects of mechanical seals, Pumps, 24(7), 209, 1968. 10. Anon., Dynamic Sealing —Theory and Practice, Koppers Company, Inc., Baltimore, Md., 1958. 11. Anon., Engineer’s Handbook of Piston Rings, Seal Rings, Mechanical Shaft Seals, 8th ed., Koppers Com- pany, Inc., Baltimore, Md., 1968. 12. Stein, P. C., Runners for circumferential seals — requirements and performance, Lubr. Eng., 36(8), 475, 1980. 13. Ruthenberg, M. L., Mating materials and environmental combinations for specific contact and clearance type seals, Lubr. Eng., 29(2), 58, 1973. 14. Wheelock, E. A., High pressure radial lip seals for rotary and recriprocating applications, Lubr, Eng., 37(6), 332, 1981. 15. Weinand, L. H., Helixseal — a practical hydrodynamic radial lip seal, ASME Trans. J. Lubr. Technol., 90(2), 433, 1968. 16. Taylor, E. D., Birotational seal designs, Lubr. Eng., 29(10), 454, 1973. 17. Horve, L. A., Reducing Operating Temperatures of Elastomeric Sealing Lips, SAE Int. Automotive Eng. Congr., SAE Paper No. 730050, January 8 to 12, 1973. Volume II 621 FIGURE 38. Viscoseal. 581-622 4/10/06 6:03 PM Page 621 Copyright © 1983 CRC Press LLC [...]...581- 622 4/10/06 622 6:03 PM Page 622 CRC Handbook of Lubrication 18 Brink, R V., The working life of a seal, Lubr Eng., 26 (10), 375, 1970 19 Schnurle, F and Upper, G., Influence of Hydrodynamics on the Performance of Radial Lip Seals, No 73AM-9B -2, American Society of Lubrication Engineers, Washington, D.C., 1973, preprint 20 Upper, G., Temperature of sealing lips, Proc 4th Int... 568, Society of Automotive Engineers, Warrendale, Pa 26 SAE, Gland Design, Aerospace Recommended Practices ARP 123 1; ARP 123 2; ARP 123 3; and ARP 123 4, Society of Automotive Engineers, Warrendale, Pa 27 Hoyle, R., How to select and use mechanical packings, Chem Eng., 103, 1978 28 Anon., Fluid Sealing, 3rd ed., George Angus and Company, Ltd., Northumberland, England, 1965 29 Anon., O-Ring Handbook, Publ... preprint 21 Dreger, D R., Ed., Materials reference issue III and IV, Mach Design, 52( 8), 1980 22 Ostmo, O., How to select shaft seal materials, Lubr Eng., 29 (6), 24 0, 1973 23 Seneczko, M., Ed., Mechanical drives reference issue III, Mach Design, 52( 14), 1980 24 Jackowski, R A., Elastomeric lip seals, Proc DOE/ASME/ASLE Seals Education Workshop, Session 9, Atlanta, Ga., October 8 to 10, 1979 25 SAE,... 13(4), 311, 1970 Copyright © 1983 CRC Press LLC Volume II 623 WEAR RESISTANT COATINGS AND SURFACE TREATMENTS S Frank Murray INTRODUCTION When it is necessary to upgrade the sliding characteristics and wear resistance of metal surfaces, coatings can often be used effectively without sacrificing any of the bulk property requirements of the substrate material In addition, the use of coatings may often provide... fine clearance with and without relative motion of the boundaries, ASME Trans., 77 (11) , 129 1, 1955 36 Egli, A., The leakage of steam through labyrinth seals, ASME Trans., 57, 115 , 1935 37 Moskowitz, R., Dynamic sealing with magnetic fluids, ASLE Trans., 18 (2) , 135, 1975 38 Stair, W K and Hale, R H., The turbulent viscoseal — theory and experiment, Paper H2, presented at 3rd Int Conf Fluid Sealing, BHRA... material and production costs The objective of this chapter is to present an overview of current practices on the use of coatings for tribological applications FACTORS TO BE CONSIDERED IN SELECTING COATINGS A wide spectrum of surface coatings or modifications are available.1 ,2 These range from soft, low friction, solid lubricant films and polymers to a number of very hard coatings Table 1 shows typical... thickness coinciding with this point of maximum shear could result in separation between the coating and the substrate The literature indicates that abrasive particles or asperities must have an angle of attack of about 80 to 120 ° to cut the surface For this reason, two-body abrasion with fixed asperities will generally cause much more wear than the three-body mode When loose particles are trapped between... 52 Rc Similarly, in the shop File Copyright © 1983 CRC Press LLC Volume II 627 FIGURE 2 Relationship between Mohs hardness number and indentation hardness (From Tabor, D., Proc Phys Soc (London), 67(3B), 24 9, 1957 With permission.) Two-Body Abrasion 1 2 3 4 5 Improve the surface texture, preferably by techniques which do not produce sharp asperities Reduce the loads Consider elastomeric coatings — particularly... Use careful run-in at light loads to wear off asperities before applying full load If severe impact loads are also encountered, select materials for fatigue resistance (toughness), with abrasion resistance as a secondary consideration Three-Body Abrasion 1 2 3 4 Prevent entry of particles by seals Provide grooves, pockets, or soft areas in surfaces to trap particles For lubricated systems, use filtration... or separators Design lubricant systems and grooves to promote flushing of debris Adhesive Wear When two surfaces are brought into contact, peaks or asperities deform plastically until the real area of contact is just sufficient to support the load elastically At these asperity Copyright © 1983 CRC Press LLC 628 CRC Handbook of Lubrication contacts, strong adhesion can occur When one surface slides . experiments on the turbulent viscoseal, ASLE Trans., 13(4), 311, 1970. 622 CRC Handbook of Lubrication 581- 622 4/10/06 6:03 PM Page 622 Copyright © 1983 CRC Press LLC WEAR RESISTANTCOATINGS AND. Society of Automotive Engineers, Warrendale, Pa. 26 . SAE, Gland Design, Aerospace Recommended Practices ARP 123 1; ARP 123 2; ARP 123 3; and ARP 123 4, Society of Automotive Engineers, Warrendale, Pa. 27 January 8 to 12, 1973. Volume II 621 FIGURE 38. Viscoseal. 581- 622 4/10/06 6:03 PM Page 621 Copyright © 1983 CRC Press LLC 18. Brink, R. V., The working life of a seal, Lubr. Eng., 26 (10), 375,