Friction and Lubrication in Mechanical Design Episode 2 Part 11 pdf

Surface Coating 483 If this equation is used for a simplified model, then we have the following equation to describe the normal stress in the diamond, interface coating, and substrate: 1

Effects of Varying Parameters

The effects of varying the sliding velocity of the solid, U, and the width of contact, I, are examined in this section Because the temperature rise in an uncoated substrate is inversely proportional to the square root of the sliding

velocity, AT oc l / m , it is expected that the temperature rise for the multi-

layered case will follow suit Figure 12.9 illustrates this case The para-

meters, except for U, are the same as in Fig 12.8 As would be expected,

the magnitude of the temperature rise for the substrate and layers dropped with the increase in sliding velocity

The effect of increasing I , the width of contact, is now considered Because I follows the same inverse relationships as U for the unlayered

substrate, AT a 1/&, it is expected that the magnitude of the temperature

rise will decrease Figure 12.10 illustrates this case The parameters, except for I, are the same as in Fig 12.8 As would be expected, the magnitude of the temperature rise for the substrate and layers dropped with the increase in contact width

Figure 12.10 Temperature rise ("C) in substrate, interface layer, and diamond

48 I

482 Chapter I2

12.8 THERMAL STRESS CONSIDERATIONS

In this section, simplified equations are developed for predicting the magni- tude of the thermal stresses in a multilayered coating The thermal stress in the substrate will be combined with the contact stress to determine a max- imum stress value for calculating a life debit due to thermal fatigue

12.8.1 Thermal Stress Relationships

Nominal stress relationships for design purposes developed in the following sections refer to the diagram in Fig 12.1 1 This figure defines the variables used in predicting normal and shear thermal stresses for the case of a multi- layer semi-infinite substrate moving under a stationary heat source with a Hertzian distribution

Surface Coating 483

If this equation is used for a simplified model, then we have the following equation to describe the normal stress in the diamond, interface coating, and substrate:

(1 2.8)

where T d , TK, and ry.y are the temperature differentials between each of the

layers and its substrate Note that Eq (12.8) is a nominal relationship for design approximation only

12.8.3 Shear Stresses

Shear stress can be determined by dividing the shear force, F,, by the shear

area, A,y The shear force can be approximated by the difference in normal stresses between two layers times the cross-sectional area of

Fy = (02 - ol)A, The shear stress can now be written as

( 12.9) Now, referring to Fig 12.1 1, if we substitute A, = / I , and A,y = /m, we have:

Figures 12.12 and 12.13 show the calculated nominal thermal stresses and interface shear stresses for the examples considered in the previous sections using equal thickness layers of silicon nitride and diamond on stainless steel

12.8.4 Life Improvement Due to Surface Coating

The effect of thermal stress on the life of a stainless steel substrate is con- sidered with and without protective coating In this example, coating layers

Normal stress (Pa) for various thicknesses (mm) of diamond and

Substrate is AISI 304 stainless steel and interface layer is silicon

h ' I/ = hd = h , , oI = ad, a2 = aq and a3 = a.s.s ( U = 12.7m/sec,

Figure 12.13 Shear stress (Pa) for various thicknesses (mm) of diamond and

interface layer Substrate is AISI 304 stainless steel and interface layer is silicon

484

Surface Coating 485

of diamond and silicon nitride of equal thickness are used as in the previous case The Hertzian contact stress is combined with the thermal stress using the Von Mises distortion energy theory for predicting the relative surface damage The results for different coating thicknesses are given in Fig 12.14

as an illustration It can be seen from the figure that considerable improve- ment in life can be expected as a result of the coating The improvement tends towards an asymptotic value for relatively thick layers

Figure 12.14 Life improvement in cycles versus thickness of diamond film in

REFERENCES

York, NY, 1991

Coating: Development, Properties, and Application, Davis, R (Ed.), Park Ridge, NJ, Noyes Publications, 1993, pp 1-30

Spear, K., and Dismukes, J (Eds), Synthetic Diamond: Emerging CVD Science

Field, J (Ed.), The Properties of Natural and Synthetic Diamond, Academic Press/Harcourt Brace Jovanovich, London, England, 1992

Singh, R., Private communications, Dept of Material Science, Univ of Florida-GainesviIle

Busch, J., and Dismukes, J., “A Comparative Assessment of CVD Diamond Manufacturing Technology and Economics,” Synthetic Diamond: Emerging CVD Science and Technology, Spear, K and Dismukes, J (Eds), John Wiley

& Sons, New York, NY, 1994, pp 581624

Moustakas, T., “Growth of Diamond by CVD Methods and Effects of Process Parameters,” Synthetic Diamond: Emerging CVD Science and Technology,

Coated Sliding Surfaces,” Thin Films in Tribology, Dowson, D., et al (Eds), Elsevier Science Publishers B.V., Amsterdam, The Netherlands, 1993, pp 399-

407

Selection,” Thin Films in Tribology, Dowson, D., et al (Eds), Elsevier Science Publishers B.V., Amsterdam, The Netherlands, 1993, pp 429-439

Rashid, M., and Seireg, A., “Heat Partition and Transient Temperature

Relationships and Numerical Results,” ASME J Tribol., 1986, pp 102-107

Mort, J., “Diamond and Diamond-like Coatings,” Mater Des., June 1990, Vol Rickerby, D S., and Matthews, A., Advanced Surface Coatings: A Handbook of Surface Engineering, Blackie and Son, New York, NY, 1991

Sander, H., and Petersohn, D., “Friction and Wear Behavior of PVD-coated Tribosystems,” Thin Films in Tribology, Dowson, D et al (Eds), Elsevier Science Publishers B.V., Amsterdam, The Netherlands, 1993, pp 483493 11(3), pp 115-121

Surface Coating 48 7

Stafford, K N., Subramanian, C., and Wilkes, T P., “Characterization and Quality

Applications, Dotta, P K et al (Eds), Royal Society of Chemistry, Cambridge,

England, 1992

Coated Cutting Tools

Anon, “Cutting Tools as Good as gold,” Metalwork Prod., July 1983, Vol 127(7), Bhat, D G., and Woerner, P F., “Coatings for Cutting Tools,” J Metals, Feb 1986,

Bollier, R D., “Recoating Enhance Resharpening,” Mod Mach Shop, March 1986,

Hale, T., and Graham, D., “How Effective Are the Carbide Coatings?” Aust Mach

Hatschek, R L., “Coatings: Revolution in HSS Tools,” Am Machin., March 1983,

Hewitt, W R., and Heminover, D., “TiN Coating Benefits Apply to Solid Carbide

Tools Too,” Cutting Tool Eng., Jan.-Feb 1986, Vol 36(1-2), pp 17-18

Jackson, D., “Coatings: Key Factor in Cutting Tool Performance,” Mach Tool Blue

Applications,” Indust Heat., Jan 1986, Vol 53( l), pp 20-22

Schintlmeister, W., Wallgram, W., Kanz, J., and Gigl, K., “Cutting Tools Materials

Coated by Chemical Vapor Deposition,” Wear, Dec 1984, Vol lOO(1-3), pp Walsh, P., and Bell, D C., “Recoatingr A Viable Option of TIN Coating for Special

Wick, C., “Coated Carbide Tools Enhance Performance,” Manu Eng., March 1987,

Wick, C., “HSS Cutting Tools Gain a Productivity Edge,” Manuf Eng., May 1987, Zichichi, C., “Tool Coatings: Trends and Perspectives,” Carbide Tool J., Jan.-Feb

pp 129-144

Vol 38(2), pp 68-69

Vol 58(10), pp 76-81

Titanium Nitride Process,” Indust Heat., Vol 53(9), pp 18-20

Prod Eng., April 1984, Vol 37(4), pp 17-19

Another experimental procedure is discussed which can be used to investigate the oil film pressure generated by a slider with different geome- tries undergoing a reciprocating motion at a predetermined distance from a flat surface

The last two sections describe experimental techniques which can be used to study the effect of the lubricant properties on surface temperature and wear in sliding contacts and the effect of repeated thermal shock on the fatigue life of high-carbon steels

13.1 FRICTIONAL INTERFACE BEHAVIOR UNDER SlNUSOlDAL FORCE EXCITATION

This section describes an experimental technique developed by Seireg and Weiter [l] for studying the vibratory behavior of a ball supported between

two frictional joints The setup which is utilized in this investigation for evaluating the “break away” coefficient of friction under sinusoidal tangen- tial forces is also useful in determining the ball response and the energy dissipated per cycle under excitations of different amplitudes and frequen-

488

Some Experimental Studies 489

cies Wear and lubrication studies can be readily performed on different contact conditions under sinusoidal tangential forces with frequencies ranging from zero to 2000 Hz and amplitudes from zero to the value neces-

sary to cause gross slip The main difference between the proposed techni- que and previous methods is that the tangential force (rather than the displacement) is sinusoidal and remains as such up to the “break away” value

The effect of an oscillating tangential force on the contact surfaces of elastic bodies has been subject to considerable interest in recent years Several valuable contributions are available in the literature Mindlin [2] extended the classical Hertz theory of contact to include the effect of an increasing tangential force with the normal force unchanged He predicted that slip would occur at the edges of the contact area and progress inwards

as the tangential force increases This slip would occur only on annular ring surfaces At any point on the contact surface where slip has just taken place, the tangential component of traction has the same sense as that of the slip, and its magnitude is equal to the product of a constant coefficient of friction and the normal component of the pressure at that point The tractions on and the displacements of the portion of the contact surface where no slip occurs are obtained from the solution of the boundary value problem Expressions for calculating the relative tangential displacement of distant points on opposite sides of the contact due to a tangential force smaller or

equal to that necessary for gross slip are given in Chapter 3 The theory was

further extended to calculate the displacement due to an oscillating tangen- tial force within the region of no gross slip The result is a hysteresis loop and the energy dissipation for the cycle due to friction can readily be calcu- lated Mindlin et al [3] found from experiments on polished crown glass

lenses that the area of the loop at low loads varied as the square of the displacement, whereas the theory predicts a cube law The agreement with the theory was good for large displacements The oscillating force in their test was obtained by utilizing a hollow cylinder of barium titanate for the driving transducer, which is essentially a displacement generator producing sinusoidal tangential displacement The force was measured by a disk of barium titanate cemented between the driving transducer and the sphere Johnson [4] utilized a torsional pendulum to apply the tangential force on three unlubricated hard steel balls on hard steel flats under a range of normal loads Johnson measured the displacements due to static and oscil- lating tangential forces within the no-gross-slip region His findings were in general agreement with the previous work Goodman and Bowie [5] used an apparatus similar to that of Ref 3 to study the damping effects at the

contacts of a 1/2 in diameter stainless steel sphere pressed between two 1/2 in square by 1/4 in thick stainless steel plates The dynamic hysteresis

490 Chapter 13

loops determined in their tests have been shown to conform to the shape predicted by Mindlin’s theory Their results of a dimensionless energy dis- sipation versus the ratio of peak-to-peak displacement at gross slip were in fair agreement with the theory

Klint [6] studied the effects of oscillating tangential forces within the

region of no gross slip on cylindrical specimens in contact A horizontal test

cylinder of 1/8 in radius is attached to the piston of a hydraulic cylinder and

is forced by the oil pressure against a vertical test cylinder attached to the table of a shaker producing smooth sinusoidal movements The hydraulic cylinder and consequently the horizontal test cylinder, although spring mounted on the shake table, are essentially fixed in space due to the vibra- tory characteristic of their support

A barium titanate force gage was used between the test specimen and

the shake table The tangential compliance and the energy dissipation per cycle were studied for different combinations of materials

A region within the no-gross-slip region was found where the displace-

ments are primarily elastic and was defined by the “limit of elastic behavior’ The coefficient of friction was calculated from the friction force represented

by the flat portion of the force-time relation Wear and surface damage conditions were also investigated in the test

In all the previous experimental procedures, the oscillating tangential force was provided by applying sinusoidal relative tangential displacements

to the bodies in contact The force wave forms appear sinusoidal at dis- placements well below gross slip and then progressively change toward waves with flat tops as the peak displacement is increased

The investigation described in this section was, therefore, planned to provide a sinusoidally changing tangential force with amplitudes up to the gross slip force, and to study its effects on a sphere pressed between two flat surfaces by a constant normal force With such a system, it would be possible to study the motion of the ball as a mass supported by a non- conservative hysteretic spring and subjected to sinusoidal excitations (refer

to Fig 13.1)

13.1.1 Experimental Setup

The apparatus is illustrated diagrammatically in Fig 13.2 The main test fixture consists of a 1: in ball (a) supported between the flat surfaces of two

cylindrical pins 0.572 in in diameter One of the pins (b) can be fixed rigidly

to the aluminum frame (c) while the other pin (d) acts as a piston in a brass air cylinder (e) attached to the frame to provide the normal force The air

cylinder pressure is controlled by a pressure regulator (f) connected to a 150

psi air supply When the ball is in place, the pins extend 1/16 in from the

Some Experimental Studies 491

Figure 13.1

tory system

Diagrammatic representation of (a) forces on the ball; (b) the vibra-

frame in order to insure maximum rigidity The test fixture is rigidly

fastened to the table (8) of a 50 lbf, 0-2000 Hz electromagnetic shaker

A differential transformer type displacement transducer (h) is rigidly

mounted on the frame with the movable core in contact with the ball and exerting a 12 gf preload The transducer excitation and amplification is

carrier AmplHior