Handbook of Lubrication Part 16 ppsx

Packing Seals Mechanical shaft packings include compression packing, automatic or lip packing, andsqueeze packing.. Leakage from a compression packing will be approximately 5 to 100 time

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 purposeseals

Packing Seals

Mechanical shaft packings include compression packing, automatic or lip packing, andsqueeze packing Compression packings are a pliable material compressed between the throatand gland of a stuffing box for reciprocating, oscillating, and rotating applications Leakage

in dynamic applications is usually on the order of 50 to 500 m/hr, but may be essentiallyzero in semistatic valve stem applications Automatic packings utilize a flexible lip energized

by the contained fluid pressure Employed primarily for reciprocating applications, heatdissipation 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 matic devices.25,26 Squeeze packings are most frequently used in reciprocating service or inlow-speed oscillating applications such as valve stems Rotary applications are recommendedonly under well-lubricated low speed conditions, 1.75 to 4 m/sec (350 to 800 ft/min) None

pneu-of these packing devices are bearings Side loads due to out-pneu-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 commonfluid seal It consists of a number of deformable packing rings or a long rope-like materialspiral wrapped around the shaft or rod, compressed by the gland to seal against the housingbore and shaft Leakage on the order of 0.01 m/hr/m/kPa (0.0018 m/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 andfriction loss will be about three times greater Compression packing has the advantage ofbeing replaceable without disassembly of equipment and a gradual leakage increase usually

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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.

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 volumefrequently swing total cost in favor of mechanical seals

Compression packings are used extensively in rotary applications such as pumps up toabout 15 m/sec for pressures up to 1000 kPa (145 psi) and valve stems under semistaticconditions up to 34,500 kPa (5000 psi) Compression packing are sometimes used for sealingreciprocating shafts but they have the disadvantage of high friction

Figure 19 shows representative designs and the most frequently used materials 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

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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

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 minimizegland 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 slightlyoversize to provide good housing bore fit A typical packing set may use very dense “anti-extrusion” rings at the throat and gland with intermediate rings graded from soft near thethroat to hard near the gland.27

A lantern ring, Figure 18, is frequently used in compression packings for rotary cations, especially at high pressures and temperatures The lantern ring has an H cross sectionand is made of rigid material such as brass, aluminum, stainless steel, or PTFE The ring

appli-is adjacent to openings in the stuffing box wall for injecting coolants or lubricants, and adischarge can be provided on the opposite side of the housing The lantern ring can also beused to (1) introduce fluid from pump discharge when pump suction is subatmospheric toprevent 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 thepacking 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 highpressure 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 unpressurizedreturn stroke are lower than on the pressure stroke and produce a “breathing” action thathelps lubricate the seals The friction of automatic packing is approximately proportional topressure up to about 7000 kPa (1015 psi) Above this, the rate of friction increase withpressure decreases and becomes quite small at about 14,000 kPa (2030 psi).28 Automaticpackings 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 extrudethe cup packing against the cylinder wall, which causes high friction, wear, and reducedsealing 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 arefrequently used because of space limitations Leather continues to be much used for flangepacking 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 intandem U-rings are chiefly employed as piston seals but can be arranged in glands Indouble-acting piston seals, the U-ring must be used heel-to-heel A lip-to-lip arrangementwill create a pressure trap and cause rapid seal wear and failure Homogeneous U-rings inShore A hardness of 70 can be used up to about 10,000 kPa (1450 psi) in precision machinedparts 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 reinforced rings should be used U-rings with metal-reinforced bases have been used up to35,000 kPa (5100 psi) Some proprietary U-ring designs having long thick-walled staticsealing lips can be installed with enough interference to make pedestal rings unnecessary

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cut, the joints spaced at 120°, to simplify replacement without machine disassembly rings are available in leather, homogeneous elastomers, fabric-reinforced elastomers, andPTFE Split rings are usually fabric reinforced Homogeneous rings are used up to about20,000 kPa (2900 psi) At pressures around 35,000 kPa (5100 psi), homogeneous rings can

V-be mixed with leather or PTFE rings, or a combination of different hardness rings can V-beused with softer, more leak-tight rings placed nearest the high pressure At pressures above45,000 kPa (6500 psi), endless fabric-reinforced elastomer or PTFE rings are common, andthin metal separators frequently support each pressure ring V-rings can be used as pistonseals but are more commonly used in rod seal glands V-rings can be designed to withstandalmost 45,000 kPa (6500 psi) per ring, but this practice results in poor seal life Three ringsare usually the fewest employed even at modest pressures At 35,000 kPa (5100 psi), atypical packing set would have five or six rings The male and female support rings areusually made from the same material as the pressure rings when used at low pressures, lessthan 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 informationprovided by the packing manufacturer A problem common to lip-type automatic packings

is extrusion due to high pressure and excess clearance Metal surfaces in sliding contactwith automatic packing should be finished to 0.2 to 0.4 µm (8 to 16 µin.) Finish shouldnot 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,25and from over a dozen elastomers with hardness ranging from 10 to 100 Shore A.21Theseseals, Figure 23, are low in cost, require minimum space, are easy to install, require noadjustment, 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 forindustrial, aerospace, and military applications Squeeze rings, though simple in form, aremade with closely held diametral and cross section tolerances To ensure long life andeffective sealing, recommended groove dimensions, surface finishes, and diametral clear-ances must be carefully followed

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FIGURE 22 V-ring automatic gland seal.

diameter slightly smaller than the O-ring OD and the groove diameter is slightly smallerthan the O-ring ID With changes in pressure and direction, a momentary leak occurs asthe ring moves from one side of the groove to the other Since this design is primarily forlow-pressure pneumatic service, about 1380 kPa (200 psi), this slight leakage is generallyacceptable 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 Withproper 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 about5% 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 mostelastomers if heated by friction while under tensile stress will contract This contraction,the Gow-Joule effect, causes further contact load, increased friction and temperature, andrapid failure O-rings and other squeeze packings are made from a large number of elastomers

in hardnesses from about 55 to 90 Shore A A standard O-ring with a hardness of 60 willseal 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), requirebackup 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 definitesealing 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,the sealing ring interface is an ordinary mechanical face seal with a fluid film bearinggeometry added to give positive separation of the surfaces The self-acting lift pads havepockets about 10 to 25 µm (0.0005 to 0.001 m) deep and pocket-to-land width ratios in

the circumferential direction of about 2:1 Axial and radial grooves keep pressure the samearound 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 separatethe primary seal ring and rotating seat

612 CRC 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.

The pressure drop and leakage occur across the sealing dam of the sealing ring The fluidfilm bearing also contributes high film stiffness such that the seal ring can dynamically trackseal seat motion This is especially important in high-speed applications where runout couldnot otherwise be tolertated A spiral groove pattern can be applied on the seal face to operate in

a manner similar to the lift pads.31with a wide radial face, pumping action of the spiral groovescan result in zero net leakage under ideal conditions

Hydrosatic Seals

There are two kinds of hydrosatic close clearance seals: self activated and externallypressurized 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 separationdecreaes 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 highpressure 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 sealingtoxic fluids if the buffer fluid is not compatible with the sealed fluid, a more complex sealsystem 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

614 CRC 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.

ments and still behave as a close clearance seal Multiple short rings can be staged for bettersealing and to accommodate shaft misalignment In high-temperature applications, thermalexpansion of the bushing must match that of the shaft.

The basic mass flow equations for incompressible constant area parallel flow34areLaminar

(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 = hm(1 + cos θ), where θis reckonedfrom the position at which h = hminimum, and  = e/hm, Equation 13 for laminar flowbecomes:

(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, hm = mean filmthickness, 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.

FIXED-GEOMETRY CLEARANCE SEALS

Buffered Bushing Seal

Bushing seals depend on small clearances between relatively moving surfaces and arecommonly used to limit leakage of liquids They are frequently used as shown in Figure 32with process fluid leakage being prevented by a reverse leak of buffer fluid To minimizeingress of buffer fluid, the primary bushing pressure differential, (pb– Pp), 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 andstationary housing The labyrinth seal has a long history and is widely used to minimizesteam 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,35or asturbulent pipe flow The actual process lies somewhere between Using the former approach,Egli36 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, ρ1

= entrance fluid density, and p1 = entrance fluid pressure.

SEALS USING SPECIALIZED CONTROL OF FLUID

A typical gap is 0.76 mm (30 mil): small enough to prevent extrusion of a solid sodium

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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.)

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.

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 Conf Fluid Sealing, No 8, May 5 to 9, 1969,

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), 240, 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, Standard O-Ring Sizes, Aerospace Standards AS 568, Society of Automotive Engineers, Warrendale,

Pa.

26 SAE, Gland Design, Aerospace Recommended Practices ARP 1231; ARP 1232; ARP 1233; and ARP 1234,

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 ORD-5700, Parker Hannifin Corporation, Lexington Ky., 1977.

30 Findlay, J A., Sneck, H J., and Reilly, J A., Final Rep on Study of Dynamic and Static Seals for

Liquid Rocket Engines, Contract NAS 7-434, Phase III, NASA CR 109646, General Electric Company, January 1970.

31 Strom, T N., Ludwig, L P., Allen, G P., and Johnson, R L., Spiral groove face seal concepts;

comparison to conventional face contact seals in sealing liquid sodium (400 to 1000°F), ASME Trans J Lubr Technol., 90(2), 450, 1968.

32 Muller, H K., Hydrodynamic and Hydrostatic Face Seals, ASLE Seals Education Course, Session 9,

Houston, Tex., May 1972.

33 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.

34 Stair, W K., Basic theory of fluid sealing, Proc DOE/ASME/ASLE Seals Education Workshop, Atlanta,

Ga., October 8 to 10, 1979.

35 Tao, L H and Donovan, W F., Through-flow in concentric and eccentric annuli of fine clearance with

and without relative motion of the boundaries, ASME Trans., 77(11), 1291, 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 Fluid Engineering, Cranfield, Bedford, England, April 1967.

39 Stair, W K., Fisher, C F., Jr., and Luttrull, L H., Further experiments on the turbulent viscoseal,

ASLE Trans., 13(4), 311, 1970.

622 CRC Handbook of Lubrication

S Frank Murray

INTRODUCTIONWhen it is necessary to upgrade the sliding characteristics and wear resistance of metalsurfaces, coatings can often be used effectively without sacrificing any of the bulk propertyrequirements of the substrate material In addition, the use of coatings may often providesavings in both raw 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 fromsoft, low friction, solid lubricant films and polymers to a number of very hard coatings

Table 1 shows typical examples, classified according to the application process When coatingsare being applied by processes such as electroplating, thermal spraying, sputtering, etc., thenumber of possible substrate/coating combinations is very large In contrast, chemical con-version and diffusion treatments are generally confined to specific classes of alloys

A detailed breakdown of various lubrication, speed, load, substrate and coating factorsinvolved in choosing a wear-resistant coating has been prepared by Czichos.3 While somematerial combinations can run dry if the operating conditions are not too severe, the greatmajority would be much more effective with some form of lubrication, even with low-viscosity fluids such as fuel or water Solid lubricant films can also provide satisfactory life

in many applications

An ideal bearing material combination would be two hard, smooth bearing surfaces,perfectly aligned with no edge contacts However, cost and fabrication problems with such

a precise system restrict its use to a very few premium applications An alternative approach

is to make one of the two surfaces considerably softer so that it can flow plastically underload, The following table shows the approximate order in which a few typical soft bearingalloys will conform and achieve fluid film lubrication:

The key appears to be the ability to develop a better surface finish and conforming geometry

in the shortest possible time

Since most soft bearing alloys have limited structural strength and fatigue resistance, theyare generally used as thin overlays on backings of steel, bronze, or aluminum alloys Asmany as two or three layers of different alloys may be applied — each serving a differentpurpose Application methods include casting, sintering, or electroplating of the individuallayers

Recent advances in polymer technology, particularly with the polyamide-imide, polyimide

Abrasive Wear

Abrasive wear is caused by penetration and cutting of a surface.5Wear caused by sharp

asperities on one surface removing material from an opposing surface is classified as body abrasion Examples are a file shaping a metal surface, chunks of minerals sliding down

two-a mettwo-al chute, or two-a rough mettwo-al surftwo-ace sliding two-agtwo-ainst two-another mettwo-al Wetwo-ar ctwo-aused by foreign

matter trapped between two moving surfaces is termed three-body abrasion This occurs

when particles are trapped in a bearing clearance or when mineral particles are being reduced

by ball milling

The amount of abrasive wear that can be tolerated varies widely In a hydrodynamic gasbearing a single scratch might cause rapid failure On the other hand, wear of mils per hourmight be tolerable in minerals handling equipment While abrasive wear is generally as-sociated with sliding, a hard particle trapped between two rolling surfaces could produce apit which would then initiate a fatigue spall.6 One note of caution: in rolling contacts thepoint of maximum shear is at some finite depth below the surface A hard coating thicknesscoinciding with this point of maximum shear could result in separation between the coatingand 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 asperitieswill generally cause much more wear than the three-body mode When loose particles aretrapped between surfaces, only a small percentage actually cut metal The rest simply ploughthrough the surfaces or roll through the loaded contact area

The volume of material removed by abrasive wear increases almost linearly with loadand sliding distance for both two-body and three-body abrasion Thus:

For applications where abrasion, impact and shock are severe, as in mining and earthmoving, a tough material is needed with high fatigue resistance.9 Austenitic manganesesteels are widely used for this type of application.10Although their hardness is only about

200 Bhn after they have been heat treated to improve toughness, these steels readily harden when they are deformed and can develop case hardnesses of 450 to as high as 550Bhn They can be used as solid members, replaceable wear strips or as welded overlays.Richardson11 showed that the hardness of surfaces must be at least half the hardness ofthe abrasive for any benefit in wear resistance Hardening the surfaces more than 1.3 timesthe abrasive hardness gave no further improvement Tabor12 showed that a metal surface ofindentation hardness Hs will be scratched by a point of hardness Hp if Hp is greater than

work-or equal to 1.2 Hs Thus, fwork-or the two-body abrasion mode, an asperity on a steel surface