Friction and Lubrication in Mechanical Design Episode 2 Part 11 pdf

480 250 200 G a E 100 k50 L -8 150 5 0) P E 0 Chapter 12 0 2 4 6 8 10 12 14 16 h, x 10' Figure 12.8 Temperature rise ("C) in substrate, interface layer, and diamond layer for diamond on AISI 304 stainless steel with a silicon nitride interface layer. (U = 12.7 m/sec, I = 0.25 mm.) 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. 120 110 100 90 o^ 80 O- 0) 70 cn 60 50 a Ot -10 1 1 -20 1 I I I I I I I 1 6 8 10 12 14 16 0 2 4 ho, x 103 Figure 12.9 Temperature rsie ("C) in substrate, interface layer, and diamond layer for diamond on AISI 304 stainless steel with a silicon nitride interface layer (velocity increase with length of contact constant). (U = 50.8 m/sec, 1 = 0.125 mm.) 100 1 I I I I I I I 1 90 n 80 Q) cn E 70 3 1 60 0) P E 50 E 8! 40 30 __ 0 2 4 6 8 10 12 14 16 h,,, x 10' Figure 12.10 Temperature rise ("C) in substrate, interface layer, and diamond layer for diamond on AISI 304 stainless steel with a silicon nitride interface layer (velocity constant with length of contact increase). (U = 12.7 m/sec, I = 1.375 mm.) 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. 12.8.2 Normal Stresses The “normal” thermal stress in an axial beam built in at both ends is proportional to the increase in temperature and can be expected as: t~ = EcYAT (1 2.7) Figure 12.1 1 Model used in calculating thermal stress in multilayer coatings. 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 Td, 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: h I t = (a2 - 01) - ( 12.10) and can write the following simplified equations for the shear stress between the diamond and interface layers, and the interface layer and the substrate: h. , I tyis = (ass - ag) rJ (12.1 1) 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 7x1 0' 6x1 0' 5x10' 8 4x10' v) 3 3x10' 0 5 3 E 2x10' 1 xl 0' 0 0 Figure 1 2.1 2 interface layer. nitride. Note: 1 = 0.25mm.) 2 4 6 8 10 12 14 16 h x103 0. 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, A a a Y t 5x1 0' 1 1 1 n. 1 -nu- $ -1~10' 0) L: cn -a1 0' -a1 0' I - 0 2 4 6 8 10 12 14 16 ho x103 . 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 nitride. Note: h, = h,l = h,, tl = rclf, and r2 = rif:ss. (U = 12.7m/sec, 1 = 0.25mm.) 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. 10’ 1 OS 1 0’ 0 2 4 6 8 10 12 14 16 h x10’ 0” Figure 12.14 Life improvement in cycles versus thickness of diamond film in mm. Substrate is AISI 304 stainless steel and interface layer is silicon nitride. Note: hjf = hd = h, and contact stress level is 1000 MPa. (U = 12.7m/sec, 1 = 0.25mm.) REFERENCES 1. Bhushan, B., and Gupta, B. K., Handbook of Tribology, McGraw-Hill, New York, NY, 1991. 2. Sherbiney, M. A., and Halling, J., “Friction and Wear of Ion-Plated Soft Metallic Films,” Wear, 1977, Vol. 45, pp. 21 1-220. Chapter I2 486 3. 4. 5. 6. 7. 8. 9. 10. 11. Yoder, M., “Diamond Properties and Applications,” Diamond Films and 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 and Technology, John Wiley & Sons, New York, NY, 1994, p. 663. 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, Spear, K. and Dismukes, J. (eds), John Wiley & Sons, New York, NY, 1994, Holmberg, K., Ronkainen, H., and Matthews, A., “Wear Mechanisms of Coated Sliding Surfaces,” Thin Films in Tribology, Dowson, D., et al. (Eds), Elsevier Science Publishers B.V., Amsterdam, The Netherlands, 1993, pp. 399- 407. Matthews, A., Holmberg, K., and Franklin, S., “A Methodology for Coating 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 Distribution in Layered Concentrated Contacts. Part I1 - Dimensionless Relationships and Numerical Results,” ASME J. Tribol., 1986, pp. 102-107. pp. 145-192. FURTHER READING Coating Bell, T., “Towards Designer Surfaces,” Met. Mater., August 1991, Vol. 7(8), pp. 478485. Gao, R., Bai, C., Xu, K., and He, J., “Bonding Strength of Films Under Cyclic Loading,” Surface Engineering Volume 11: Engineering Applications, Dotta, P. K. et al. (eds), Royal Society of Chemistry, Cambridge, England, 1992. 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 Assurance of Advanced Coatings,” Surface Engineering Volume 11: Engineering 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, Garside, B. L., “Improvements in Tools and Product Performance Through PVD 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 Bk., Vol. 81(1), pp. 62-64. Kane, G. E., “Modern Trends in Cutting Tools,” Society of Manufacturing Engineering, 1982, pp. 54-55. Kane, G. E., “Modern Trends in Cutting Tools,” Society of Manufacturing Engineers, 1982, pp. 82-87. Podop, M., “Sputter Ion Plating of Titanium Nitride Coatings for Tooling 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 Tooling Applications,” Cutting Tool Eng., Feb. 1986, Vol. 38(1), pp. 25-27. 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. Vol. 127(3), pp. 129-144. 153-1 59. Vol. 98(3), pp. 45-50. Vol. 98(5), pp. 39-42. 1986, Vol. 18(1), pp. 18-20. 13 Some Experimental Studies in Friction, Lubrication, Wear, and Thermal Shock This chapter describes a number of experimental investigations covering different aspects of tribology. The first set of experiments deals with the behavior of Hertzian frictional contacts under different types of tangential loading. In this set, unlubricated spheres pressed against flat surfaces are subjected to oscillatory loads, impulsive loads, and ramp-type loads respec- tively. 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 [...]... slip F/Fo = 1 and substitution in Eq (13.9) gives: - = 0 .26 d 0 (13.10) 60 In this case: which can be written as: pk = 5 0 7 5 ~ ~0.13 92( N)'j36; - 2 N ( d - O .26 aO) (13. 12) Some Experimenta1 Studies 503 or in the form: d= + 5 0 7 5 ~ ~0.13 92( N)''36% 0.52pkN60 2pkN (13.13) Equations (13 .11) , (13. 12) , and (13.13) are utilized for evaluation of the frictional characteristics of the joint from the experimental... 499 A steel ball (g), 1 /2 in in diameter is cemented to a fine thread and suspended as shown in Fig 13.6 to form an 87.3 in pendulum This suspension insures the fall and rebound of the ball to be in the same plane The initial and final position (after rebound) of the pendulum can be read on a graduated arc (h), giving accurate indication on the velocity of the impact ball before and after impact The velocity... alcohol and paper towels The oil is applied to the surface by wiping with a clean, predipped cloth The alcohol is applied by simply dipping the ball (after cleaning) into the alcohol and allowing it to dry in the atmosphere Throughout this process, the ball is handled using a plastic holder for each contaminating liquid The pins are prepared for each test by finishing their test surfaces with 4/9 sandpaper... 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 0.5 72 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... displacement a0 for and gross slip can therefore be obtained from the equation: 0 5 10 15 20 Velocity &Approach, V (id8) Figure 13.8 Evaluation of conditions at gross slip 25 504 Chapter 13 do 0 .26 So = -in * and the tangential force for gross slip is therefore: from which the coefficient of friction for gross slip is: F~ p, - - = 2. 27 '-2N 105 ~113 2N 60 d 0 = 0.437 N2/3 ( 1 3.14) By substituting the value... IMPULSIVE LOADING There is considerable practical interest in determining the frictional resistance under impulsive loading Gaylord and Shu [9] reported that some materials (steel on steel and titanium on steel) exhibited higher static coefficients of friction under statically applied loads than under shock loads In 0.0 I 0 I 40 I loo I I I 20 0 300 400 II 500 Shake Table Frequency, f (Hz) Figure 13.5 lubrication. .. acceleration and displacement are shown in Fig 13.3 A typical plot of the test results is shown in Fig 13.4, and the coefficients of friction are given in Table 13 .2 The lines representing the relation between the amplitude T+ of the tangential force versus the air pressure and normal force N were found to fit both the dynamic data and the static data For all the tests performed in this investigation,... calibration was done by means of a scale against the movable pin For the static friction tests, the specimens are placed in position and the pressure adjusted for a certain normal force The tangential force is applied at the lowest point of the ball in line with the displacement transducer The load is increased successively until gross slip is observed by watching the indicator on the transducer bridge unit... direct writing oscillograph The record is calibrated to give the final ball displacement a 1 pin sensitivity Accurate alignment is provided so that the impact between the two spheres is on the same axis as the displacement transducer 13 .2. 2 Test Procedure The surface preparation consisted of washing all contacting surfaces of the balls and pins with methyl alcohol The pins are placed in the frame and the... = 1 .22 8 Area oabo ( 1 3.8) Because there is no way to detect gross slip in this investigation except by means of the permanent displacement of the ball, Mindlin’s theory is again utilized to calculate this value This permanent displacement within the region of no gross slip is given by: Chapter 13 5 02 616, Figure 13.7 Dimensionless displacement of ball within the region of no gross slip (Mindlin's . Trends in Cutting Tools,” Society of Manufacturing Engineering, 19 82, pp. 54-55. Kane, G. E., “Modern Trends in Cutting Tools,” Society of Manufacturing Engineers, 19 82, pp. 82- 87. Podop,. 0.105 0. 120 0. 121 13 .2 FRICTION UNDER IMPULSIVE LOADING There is considerable practical interest in determining the frictional resis- tance under impulsive loading. Gaylord and Shu [9]. 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