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Tribological, thermal and mechanical coupling aspects of the dry sliding contact
ARTICLE IN PRESS
Tribology International 40 (2007) 834–843 www.elsevier.com/locate/triboint
Tribological, thermal and mechanical coupling aspects of the dry sliding contact$ D. Majcherczaka,, P. Dufrenoya, Y. Berthierb a
Laboratoire de Me´canique de Lille, UMR CNRS 8107, Polytech’Lille, 59655 Villeneuve d’Ascq, France Laboratoire de Me´canique des contacts et des solides, UMR CNRS 5514, INSA de LYON, 69621 Villeurbanne, France
b
Received 11 October 2005; received in revised form 10 August 2006; accepted 22 August 2006 Available online 9 October 2006
Abstract An original experimental set-up, made of two coaxial rings in relative motion, a sapphire and steel, enabled temperature measurements on both sides of the third body at the friction interface. Hot spots have been identified and temperature gradient across the third body accurately measured. Infrared camera and thermocouples have shown to be an effective tool for this research. Investigations conducted using SEM enabled detailed analysis of friction interfaces of both components, the sapphire and steel rings. Two types of third body (layers) have been identified, the compact, smooth micro-plates—where the actual contact occurs, and granular—which seem to accumulate in depressions or against material obstacles. There are also clear indications that the hot spots and depressions on steel friction surface are directly related. These areas of contact seem to be ‘shrinking’ in height after the application and complete component cooling. The investigation of the third body phenomenon and its influence on interface temperatures has direct relation to the observations made in automotive disc brakes. A thermal numerical model, which was also developed, introduced the third body as a uniform layer with energy storage and conduction. The obtained thermal gradients seem to be accurate, when compared with measurements conducted. The results are also similar to those found in literature. In addition, when only a fraction (1/1000th and 1/2000th) of the total nominal friction surface was considered to be in the actual contact, experimental temperature results were exactly within the predicted range. This indicates that the actual contact area varies during application. r 2006 Elsevier Ltd. All rights reserved. Keywords: Thermal contact; Third body; Infrared thermography; Braking
1. Introduction The mechanisms taking place in a sliding contact include several strongly coupled physical phenomena: mechanical loading (perpendicular to the surface), shear loading (in the sliding plane), heat generation, heat evacuation from the interface (heat transfer from the interface and further conduction and heat dissipation), chemical transformations, material degradations, etc. When considering sliding con$ This paper was presented at the 32nd Leeds-Lyon Symposium in Lyon, 6–9 September 2005. Corresponding author. E-mail addresses: [email protected] (D. Majcherczak), [email protected] (P. Dufrenoy), [email protected] (Y. Berthier).
0301-679X/$ - see front matter r 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.triboint.2006.08.004
tacts, it is important to understand the mechanisms of energy transformation (generation). The studies show that 95% of the mechanical energy is transformed into heat. The remaining 5% is dissipated through noise, light, gaseous emissions, etc. Since the majority of energy is transformed into heat, it is important to understand the mechanisms governing friction generation and the influence they have on the friction pair—materials in contact. When studying these mechanisms, of particular importance are the locations of heat generation, processes related to, and influencing local heat generation, i.e. interface changes such as: third layer/ body, thermal distortion, changes of friction pairs (materials) in contact, etc. Considering geometrical and loading conditions, there is no doubt that the research of sliding contact requires a multi-scale approach.
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To illustrate sliding contact, an example of friction brake will be used. The thermomechanical, dynamical and corresponding tribological phenomena will be investigated. So far, these processes have been studied separately. Thermal and thermomechanical behaviour are generally described using macroscopic models [1–5]. Interface effects at the contact surfaces (tribology) are generally investigated with microscopic or mesoscopic qualitative models, from rheological and physico-chemical points of view [6,7]. For illustration purposes, Fig. 1a shows a railway disc brake assembly during high speed application on a test bench (inertia dynamometer). Fig. 1b shows microscopic photograph of the friction interface of the brake pad after such an application. In friction brakes, the thermal phenomenon is of an eminent importance. Vehicle deceleration and stopping rely entirely on friction (sliding contact), and the process must be predictable and reliable in order to enable safe operation. It should be noted that in addition to substantial mechanical forces, friction heat generation is extremely high. In heavy duty brake application, the heat flux at the interface is of the order of MW/m2. The heat generated during braking causes temperature increase at the interface, which spreads fast through the brake components. Such severe thermal processes modify friction properties of the materials in contact, cause wear (Fig. 1b) and, on a large scale, result in component deflection. All these changes inevitably affect brake performance and life. Nowadays, the geometry of brake components tends to be more and more complex, especially in railway applications, with ventilated discs having complex vane/pillar patterns and pads made of small individual friction parts (Fig. 1a). As a result, structural behaviour becomes dependant upon very complex geometries.
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In order to reduce this complexity, a simpler experimental set-up has been developed and the contact interface studies conduced included both, thermal and SEM procedures. A thermal model is developed, which included a volumetric third body (layer) at the interface, with both energy storage and conduction. Finally, comparisons between modelling and experimental results have been conducted. 2. Braking phenomenon 2.1. Macroscopic scale The heat generated in frictional mechanisms, such as brakes and clutches, induces thermal dilatations leading to localised contact areas and hot spots. These localised effects may cause thermal damage and early failures. Anderson and Knapp [8] first proposed a classification of the hot spots observed in automotive braking systems. From experimental investigations on the rubbing surface of brake discs on a testing bench, Dufre´noy et al. [15] gave similar classification for railway brakes, based on thermal infrared measurements (Fig. 2):
Asperity type (Fig. 2-1) result from discrete asperity contacts. Temperatures rise rapidly but briefly, as flash temperatures, on very small areas of the rubbing surface [9]. Rheology of the contact surfaces and the surfaces topography are the main parameters to study this phenomenon [10,11]. Gradients on hot bands (Fig. 2-2) correspond to small contact sites appearing along a single rubbing path. This phenomenon is governed by the contact instabilities and the influence of the pad design, such as thermo elastic instabilities (TEI) [12–14].
Fig. 1. Braking: a multiscale phenomenon ((a) macroscopic [SNCF] and (b) microscopic [10]).
Fig. 2. Classification of hot spots illustrated by thermographs on railway brake discs [15].
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Hot bands (Fig. 2-3) appear as reduced contact friction areas of the pad in the radial direction. They are seen on the disc as narrow rings of high temperatures in the direction of sliding. They can move in the radial direction during braking, according to the evolution of the bearing surfaces. These phenomena, governed by the pad, can be modelled by ‘‘thermo-mechanical loops’’, taking into account the variation of the contact pressure distribution with thermal dilatations and wear of the first bodies [15]. Macroscopic hot spots (MHS) (Fig. 2-4) are large thermal gradients regularly distributed on the disc surface. They are fixed on the disc and appear as a buckling pattern of the disc. This phenomenon reduces drastically the contact surface area with high local temperatures. Recent developments have shown that they can be modelled by successive steps of thermomechanical bending of the disc under specific conditions [16]. MHS correspond to a waviness deformation of the disc exposed to high thermal loading. Regional hot spots (Fig. 2-5) are low thermal gradients on the whole surface of the disc, due to inhomogeneous cooling. Such distributions appear at the end of braking due to thermal diffusion and the non-uniformity of cooling. This phenomenon is mainly governed by the disc.
2.2. Microscopic scale The study of dry friction between rubbing surfaces is generally closely related to the debris rheology of the interface. The debris particles have often been considered as worn particles from the contact area. Numerous experimental studies emphasised the complexity of the mechanisms in sliding contact (third body separation into granular form and sheets of micro-plates, third body flows, etc.) [6,7,17,18]. Many observations showed that the accommodation of relative speed between the first bodies takes place by degradations of the contact surfaces, with cracks and superficial tribological transformations (STT), at a depth ranging from a nanometre to several micrometres. It appears that these STTs are real material flows parallel to the contact surface, resulting from very highly local pressure, temperature and surface contact shear [19]. The third body mechanical behaviour has been widely studied. Most of the published work is based on observations of sliding surfaces after contact opening and understanding of third body elaboration. Few studies (such as [20]) have been devoted to the examination of the interaction between heating areas and third body generation. An obvious difficulty when conducting such experiments is related to the simultaneous observation of the third body formation and the temperature distribution in a sliding contact. In order to address this difficulty in an efficient manner, a simpler experimental set-up has been proposed, which will be explained in detail.
3. The steel–sapphire rings experiment 3.1. Experimental set-up In order to improve the understanding of heat generation mechanisms and surface degradations, an original experimental set-up has been proposed. As shown in Fig. 3, two co-axial rings, made of sapphire and C35 steel are used. The inner and outer radii are 10 and 15 mm, respectively. More details of the set-up are given in [21]. The sapphire ring rotates around its axis, with the steel ring being fixed. The sapphire has been chosen as a material for the rotating ring because of its infrared quasi-transparency (transmission of 80% in the wave length of 3.5–5 mm). In order to obtain an axial heat flow into the rings, two thermal organic fibres insulators are fixed inside and outside the rings (Fig. 3). For temperature measurements, three thermocouples were used at steel ring surface and infrared camera was used to monitor interface temperatures through the sapphire ring. The thermocouples, inserted at the friction surface of the steel ring, are made of four layers: a Kaptons layer bonded on the steel surface, layers of each conductor (copper and Constantans) separated by an electrical insulator. The junction of the two conductors is made at the extremity of the thermocouple (Fig. 3). Three grooves are machined on the inner radius of the steel ring surface at equidistant angular positions. The grooves are curved in order to locate the thermocouple junction at the mean swept radius of the steel ring (Fig. 3). The main difficulty related to the use of an infrared camera, is the non-uniformity and variations of emissivity. A good example of necessary caution when studying thermograms has been given in [21]. Additional measurements with thermocouples have lead to the evaluation of a global emissivity coefficient, which is increasing during the tests as the third body accumulates on the fretting surface. In the general case, infrared measurements show that heating zones are much smaller than the apparent contact surface. 3.2. Results Three types of infrared thermographs may be distinguished according to the shape and the movement of the heating areas. Fig. 4 shows successive infrared thermographs for these three types (the dashed lines correspond to the inner and outer radii of the rings):
Fixed areas (Fig. 4a). The contact is localised on two main areas, with the temperature rise of the order of 200 1C. These areas are fixed from one image to the next. Since the steel ring is not rotating, it can be concluded that these areas are associated with steel contact surface. Moving areas (Fig. 4b), rotating at the same speed as the sapphire ring. There is a 2401 rotation between two successive pictures. Usually, the hot spot areas are more
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Fig. 3. Experimental set-up.
Fig. 4. Infrared measurements.
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elongated than the fixed areas and the temperature increase is lower (of the order of 150 1C). These areas have the same rotational speed as the sapphire ring, which suggest that these areas are associated with the sapphire contact surface. Low speed moving areas (Fig. 4c) with temperatures of the order of 100 1C. During some sequences, angular low speed rotation of the heating zones can be seen, generally after a period of heating (not at the beginning). Fig. 4c illustrates this angular movement of a heating area at three times of a sequence, separated by 29 revolutions between the first and the second and by 56 revolutions between the second and the third. Slight angular movement in the direction of rotation can be seen. Continuous observation of the thermographs shows discontinuous variations of movements for these areas, leading to the conclusion that this phenomenon is most associated with the corresponding third body flows.
Fig. 5 shows macroscopic observations (obtained using SEM) on the steel and sapphire contact surfaces, clearly displaying high surface damage. Numerous traces are observed on the steel contact surface caused by ploughing (Fig. 5a). The contact severity is also visible on the sapphire
contact surface by V oriented cracks centred on the mean contact radius (Fig. 5b). The black arrows indicate the sliding direction (Fig. 5a and b). Fig. 6 shows SEM pictures on the sapphire surface. Two types of the third body can be distinguished. The first one is compact and smooth—the micro-plates in the areas of actual contact. The second one is granular and seems to accumulate in hollows or against material obstacles to form compacted micro-plates. During tests with high loading conditions, heating zones were concentrated, quasi-systematically fixed in rotation. Thermographs show that theses zones appear systematically at the same place and that they are fixed during numerous revolutions. A schematic presentation of the contact area, drawn at the centre of Fig. 7 shows the location of each heating zones through the sapphire during the sequences on the contact surface. Micrographs taken after the tests are shown for each of the depressions, and their position clearly marked. Comparison between these observations and the thermographs (middle of the sequence for thermograph on the top of Fig. 7, and end of sequence for thermograph on the bottom of Fig. 7) show a correlation between the heating zones and depressions observed on the steel surface. Micrographs
Fig. 5. Contact surfaces after the test ((a) C35 steel and (b) sapphire).
Fig. 6. The two-third body types on sapphire ring surface.
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Fig. 7. Comparison between fixed heated areas and depressions on the steel surface.
clearly indicated that depressions have been developed at the steel ring surface systematically on each heating zone. In order to explain the temperature gradient at the interface, and the wear process, two observations will be made:
Non uniformities on the surface—the local areas where the third body accumulates to form compact zones creating contacts. Heat is then generated on these small areas—explaining the high temperatures reached during the tests. Due to thermal expansion, the contact zones
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are taking more contact load and generating more heat, resulting in their distortion. The micro-plates formations can also be interpreted in terms of hot spots [9]. Thermomechanical contact instabilities may appear, depending on the sliding condition—velocity and pressure, leading to nonuniform pressure and local heating. Furthermore, high temperatures reached during the test on compacted zones of the contact surface, suggest the presence of asperity or focal hot spots. Asperity hot spots may be associated with the surface roughness of the materials. Such large temperatures and pressures cause the asperities to wear out rapidly, leaving small marks on the surface. These hot spots are associated with the phenomenon of flash temperatures [22–24]. Focal hot spots are due to thermomechanical instabilities in the sense explained by Barber [12], such as banding hot spots appearing as strips distributed along the sliding surface. They have been clearly observed in the conducted experiments, and lead to more continuous heating than asperity hot spots. The local depressions on the steel surface are most probably the consequence of these focal hot spots. These raised areas are taking the load and therefore most heat is generated here, hence hot spots are being formed in these locations. They are also zones of third body compaction and shearing (such as ‘wear’) is high. Once the experiment (heat generation) stops, these areas will cool down and contract more than the surrounding—that is why ‘depressions’ are created.
These two mechanisms, of third body accumulation and focal hot spots creation, are seen as complementary effects. Focal hot spots are mainly governed by the behaviour of the first bodies (deflections, nonuniform thermal expansion, etc.), leading to third body accumulation. The third body is accumulated and compacted on surface irregularities giving local perturbations leading to thermomechanical instabilities and focal hot spots. This illustrates the necessity to consider interaction between the local thermomechanical behaviour at the interface and the global
interaction between the first bodies (friction pair components). 4. Numerical modelling 4.1. Thermal macroscopic results The thermal problem is generally separately solved by assuming thermal continuity at the interface [25]. In this work, an interface layer, modelled from tribological considerations, is introduced in the thermal problem. The aim is to better understand the mechanisms of heat generation and to give indications of the temperatures near the contact surface. The mechanical aspects are deliberately ignored at this stage. In this model, the third body layer is introduced, between the first bodies, as a continuous thin layer [25]. This layer is volumic with uniform heat generation. Contrary to the classical modelling approach, this model enables energy storage inside the third body (layer) giving thermal distribution across its thickness. The results are mainly sensitive to two parameters: thickness and conductivity of the third body. Taking parameter values from the literature [26], the thickness of 0.01 mm and conductivity of 0.07 W m 1 K 1, results show thermal gradients between the two first body surfaces in agreement with experimental observations found in literature and obtained by the authors [25]. Fig. 8 illustrates these results by showing the temperatures of disc and pad surfaces obtained by the model developed by the authors [25], for an automotive disc brake application. A very high thermal gradient is observed between the two surfaces (of the first bodies) in contact. Pad temperature is higher than disc temperature due to the third body acting as a thermal barrier, creating thermal resistance. As the pad effusivity is lower, its temperature quickly rises in the same form as the heat flux transient evolution. Temperature differences shown in Fig. 8 are very sensitive to the thickness of the third layer (body). Fig. 9
Fig. 8. Disc and pad surface temperature evolution during an automotive braking [25].
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shows the temperature distributions in the third body during braking, for the third body thicknesses of 0.01 mm (Fig. 9a) and 0.1 mm (Fig. 9b). Temperature distribution is represented at different location into the third body. In both cases, a maximal temperature is obtained inside the third body. For thin layers, maximal temperature is obtained near the pad surface because of its lower effusivity. As thickness increases, maximal temperature moves towards the middle of the third body. Results show that introducing a third body leads to higher predicted temperatures. The contact pressure has been assumed to be uniform, but it can be considered that if the areas of real contact are smaller (i.e. assuming localised contact), temperatures would be even higher. This illustrates the necessity to enrich the models by taking into account real contact surface evaluations and material degradations. 4.2. Thermal microscopic results This model has been used in the case of the two rings experimental set-up previously described (Fig. 3). A third body (layer), having a thickness of 5 mm and thermal contact conductance of 7000 W m 2 K 1 has been modelled. The value of thermal conductance is in good
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agreement with those obtained by Day [26]. The contact pressure is assumed to be uniform and thermal analysis conducted in order to predict the temperature gradient at the interface. The numerical results are compared with the experimental thermal measurements obtained in two places: on the top of the third body layer (using the infrared camera—measuring through the sapphire), and on the bottom of the third body layer (using the thermocouples at the steel surface). Fig. 10 shows the comparison of experimental and numerical temperatures. Fig. 10a shows infrared measurements, whilst Fig. 10b shows the evolution of the average temperature obtained by the thermocouples and infrared camera, as well as the average numerical modelling temperature results. For a continuous test, under low loading conditions, the temperature distribution was uniform enough; therefore good agreement with temperature values measured using thermocouples (Fig. 10b) has been achieved. If the experimental measurements are interpreted as temperatures above and under the third body layer, the experimental and numerical thermal gradients are in good agreement in terms of gradients between the two surfaces. However, predicted temperatures are lower than those obtained experimentally. This may be due to the effective contact area assumed in the model as the whole surface,
Fig. 9. Temperature distribution into the third body for two thicknesses ((a) 0.01 mm and (b) 0.1 mm).
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Fig. 10. Comparison between experiments and numerical model ((a) contact surface and (b) thermal results).
Fig. 11. Localised contact ((a) estimated contact surface and (b) comparison between numerical and experimental results).
whereas in reality it seems to be reduced, as shown in the thermograms (Fig. 10a). If contact surface is reduced to half of the apparent one, near the outer radius as shown in Fig. 10a, the numerical results would be closer to the experimental ones. During most of the experiments, the contact area was considerably reduced due to local heating. Analysis of the thermograms allows to estimate the contact surface, which for the test sequence, is ranging from 1/2000th to 1/1000th of the nominal ring friction surface (Fig. 11a). The thermal model with third body has been used in this case, modelling a transient analysis of 10 s duration. The predicted temperatures are compared with the experimentally determined temperatures (thermographs) obtained using infrared camera. Results are presented in Fig. 11b for the two values of estimated contact surface area, Surface 1 assumed to be 1/1000th of the nominal friction surface and Surface 2, equalling 1/2000th of the nominal friction area. The maximal value of temperature on the thermograms during the sequence, marked ‘‘Experiment’’, is also shown. Fig. 11b shows good agreement between the predicted and measured values in terms of transient nominal temperature change during the 10 s event (sequence). Variations of the maximum experimental temperature value are most probably the result of the variation of the
actual contact area during the sequence. The modelling values, chosen to be 1/1000th and 1/2000th of the nominal friction area, can be considered as lower and upper limits of the actual contact area, since maximal experimental temperature values vary between these two temperature curves. 5. Conclusions The developed original experimental set-up, made of two coaxial rings in relative motion, enabled thermal measurements on both sides of the third body. Hot spots have been identified and temperature gradient across the third body accurately measured. Infrared camera and thermocouples have been shown to be an effective tool for this research. Further investigations, using SEM, enabled detailed analysis of friction interfaces of both components, the sapphire and steel rings. Two types of third body layers have been identified, the compact and smooth microplates (where the actual contact occurs) and granular (which seem to accumulate in depressions or against material obstacles). When interface analyses were compared to thermal results, a very important pattern has been discovered. There are clear indications that the hot spots and
ARTICLE IN PRESS D. Majcherczak et al. / Tribology International 40 (2007) 834–843
depressions on steel friction surface are directly related. These areas of contact seem to be ‘shrinking’ in height after the application and complete component cooling. The investigation of the third body phenomenon and its influence on interface temperatures (i.e. thermal gradient across the third body layer) has direct relation to the observations made in automotive disc brakes. The thickness of third layer may very, which directly influences thermal gradients. A thermal numerical model, which was also developed, introduced the third body as a uniform layer with energy storage and conduction. The obtained thermal gradients seem to be accurate, when compared with measurements conducted. The results are also similar to those found in literature. In addition, when only a fraction (1/1000th and 1/2000th) of the total friction was considered to be in the actual contact, experimental temperature results were exactly within the predicted range. This indicates that the actual contact area varies during application. Further work will be focused on the introduction of the thermomechanical dimension in the simulations. Contact localisation has to be introduced at two complementary scales: macroscopic with the thermal dilatations of the first bodies (actual components) and microscopic with the rheology of the surfaces and the third body formation and accumulation. References [1] Missori S, Sili A. Optimizing proportions of railway brake discs by temperature transients evaluation. Proc Inst Mech Eng Part F: J Rail Rapid Transit 1988;202:91–9. [2] Newcomb TP. Temperatures reached in disc brakes. J Mech Eng Sci 1960;2(3):167–77. [3] Day AJ, Newcomb TP. The dissipation of frictional energy from interface of an annular disc brake. Proc Inst Mech Eng Part F: J Rail Rapid Transit 1998;198:201–9. [4] Kennedy FE, Ling FF. A thermal, thermoelastic and wear simulation of a high energy sliding contact problem. ASME J Lub Tech 1974;96(2):497–504. [5] Kao T, Richmond JW, Douarre A. Brake disc hot spotting and thermal judder: an experimental and finite element study. Int J Vehicle Des 2000;23:276–96. [6] Berthier Y. Experimental evidence for friction and wear modelling. Wear 1990;115:607–14. [7] Desplanques Y, Degallaix G, Copin R, Berthier Y, A tribometer for the study of materials under braking condition. In: G. Dalmaz, et al., editors. Tribology research: from model experiment to industrial
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problem. Tribology series, vol. 39. Elsevier, Amsterdam, 2001. p. 381–91. Anderson AE, Knapp RA. Hot spotting in automotive friction systems. Wear 1990;135:319–37. Parker RC, Marshall PR. The measurement of the temperature of sliding surfaces with particular reference to railway blocks. Proc IMechE 1948;158:209–29. Roussette O, Desplanques Y, Degallaix G, Gallo Y, Comportement tribologique en freinage a` haute e´nergie de garnitures en mate´riaux organiques. JFT 2001 Tribologie des Mate´riaux Organiques, Obernai, France, 2001. Bettge D, Starcevic J. Topographic properties of the contact zones of wear surfaces in disc brakes. Wear 2003;254:195–202. Barber JR. Thermoelasticity instabilities in the sliding of conforming solids. Proc R Soc A 1969;312:381–94. Barber JR, Comninou M. Thermoelastic contact problems. In: Hetnarski R, editor. Thermal stresses III. Amsterdam: NorthHolland; 1989. p. 1–106. Du S, Zagrodzki P, Barber JR, Hulbert GM. Finite element analysis of frictionally-excited thermoelastic instability. J Therm Stresses 1997;20:185–201. Dufre´noy P, Weichert D. Prediction of railway disc brake temperatures taking the bearing surface variations into account. Proc IMechE Part F: J Rail Rapid Transit 1995;209:67–76. Panier S, Dufrenoy P, Weichert D. An experimental investigation of hot spots in railway disc brakes. Wear 2004;256:764–73. Y. Berthier, Maurice Godet’s third body. In: Proceedings of the 22nd Leeds-Lyon symposium ‘‘The third body concept: interpretation of tribologica’’, Lyon, France. Tribology series 31. Amsterdam: Elsevier; 1996. p. 21–30. Descartes S, Berthier Y. Frottement et usure e´tudie´s a` partir de la rhe´ologie et des de´bits de 3e`me corps solide. Mate´riaux & Techniques 2001;1–2:3–14. Eleod A, Bailler L, Berthier Y, Torkoly T. Deformability of the near surface layer of the first body. In: Proceedings of the 29th Leeds-Lyon Symposium on Tribology, Leeds (GB). Tribology series 41. Elsevier, Amsterdam, 2002. p. 123–132. Colombie C, Berthier Y, Floquet A, Vincent L, Godet M. Fretting: load carrying capacity of wear debris. ASME J Tribol 1984;106(2): 194–201. Majcherczak D, Dufre´noy P, Berthier Y, Naı¨ t-Abdelaziz M. Experimental thermal study of a contact with third body. Wear 2006;261(5–6):467–76. Kreitlow W, Schrodter F, Matth H. Vibration and hum of disc brakes under load, SAE 850079, 1985. Jaeger JC. Moving sources of heat and the temperature at sliding contacts. Proc Roy Soc N S W 1942;76:203–24. Kuhlmann-Wilsdorf D. Temperatures at the interfacial contact spots: dependence on velocity and on role reversal of two materials in sliding contact. J Tribol 1987;109:321–9. Majcherczak D, Dufre´noy P, Naı¨ t-Abdelaziz M. Third body influence on thermal friction contact problems: application to braking. ASME J Tribol 2005;127(1):89–95. Day AJ. Energy transformation at the friction interface of a brake. PhD thesis, Loughborough University of Technology, 1983.
Tribology International 40 (2007) 834–843 www.elsevier.com/locate/triboint
Tribological, thermal and mechanical coupling aspects of the dry sliding contact$ D. Majcherczaka,, P. Dufrenoya, Y. Berthierb a
Laboratoire de Me´canique de Lille, UMR CNRS 8107, Polytech’Lille, 59655 Villeneuve d’Ascq, France Laboratoire de Me´canique des contacts et des solides, UMR CNRS 5514, INSA de LYON, 69621 Villeurbanne, France
b
Received 11 October 2005; received in revised form 10 August 2006; accepted 22 August 2006 Available online 9 October 2006
Abstract An original experimental set-up, made of two coaxial rings in relative motion, a sapphire and steel, enabled temperature measurements on both sides of the third body at the friction interface. Hot spots have been identified and temperature gradient across the third body accurately measured. Infrared camera and thermocouples have shown to be an effective tool for this research. Investigations conducted using SEM enabled detailed analysis of friction interfaces of both components, the sapphire and steel rings. Two types of third body (layers) have been identified, the compact, smooth micro-plates—where the actual contact occurs, and granular—which seem to accumulate in depressions or against material obstacles. There are also clear indications that the hot spots and depressions on steel friction surface are directly related. These areas of contact seem to be ‘shrinking’ in height after the application and complete component cooling. The investigation of the third body phenomenon and its influence on interface temperatures has direct relation to the observations made in automotive disc brakes. A thermal numerical model, which was also developed, introduced the third body as a uniform layer with energy storage and conduction. The obtained thermal gradients seem to be accurate, when compared with measurements conducted. The results are also similar to those found in literature. In addition, when only a fraction (1/1000th and 1/2000th) of the total nominal friction surface was considered to be in the actual contact, experimental temperature results were exactly within the predicted range. This indicates that the actual contact area varies during application. r 2006 Elsevier Ltd. All rights reserved. Keywords: Thermal contact; Third body; Infrared thermography; Braking
1. Introduction The mechanisms taking place in a sliding contact include several strongly coupled physical phenomena: mechanical loading (perpendicular to the surface), shear loading (in the sliding plane), heat generation, heat evacuation from the interface (heat transfer from the interface and further conduction and heat dissipation), chemical transformations, material degradations, etc. When considering sliding con$ This paper was presented at the 32nd Leeds-Lyon Symposium in Lyon, 6–9 September 2005. Corresponding author. E-mail addresses: [email protected] (D. Majcherczak), [email protected] (P. Dufrenoy), [email protected] (Y. Berthier).
0301-679X/$ - see front matter r 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.triboint.2006.08.004
tacts, it is important to understand the mechanisms of energy transformation (generation). The studies show that 95% of the mechanical energy is transformed into heat. The remaining 5% is dissipated through noise, light, gaseous emissions, etc. Since the majority of energy is transformed into heat, it is important to understand the mechanisms governing friction generation and the influence they have on the friction pair—materials in contact. When studying these mechanisms, of particular importance are the locations of heat generation, processes related to, and influencing local heat generation, i.e. interface changes such as: third layer/ body, thermal distortion, changes of friction pairs (materials) in contact, etc. Considering geometrical and loading conditions, there is no doubt that the research of sliding contact requires a multi-scale approach.
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To illustrate sliding contact, an example of friction brake will be used. The thermomechanical, dynamical and corresponding tribological phenomena will be investigated. So far, these processes have been studied separately. Thermal and thermomechanical behaviour are generally described using macroscopic models [1–5]. Interface effects at the contact surfaces (tribology) are generally investigated with microscopic or mesoscopic qualitative models, from rheological and physico-chemical points of view [6,7]. For illustration purposes, Fig. 1a shows a railway disc brake assembly during high speed application on a test bench (inertia dynamometer). Fig. 1b shows microscopic photograph of the friction interface of the brake pad after such an application. In friction brakes, the thermal phenomenon is of an eminent importance. Vehicle deceleration and stopping rely entirely on friction (sliding contact), and the process must be predictable and reliable in order to enable safe operation. It should be noted that in addition to substantial mechanical forces, friction heat generation is extremely high. In heavy duty brake application, the heat flux at the interface is of the order of MW/m2. The heat generated during braking causes temperature increase at the interface, which spreads fast through the brake components. Such severe thermal processes modify friction properties of the materials in contact, cause wear (Fig. 1b) and, on a large scale, result in component deflection. All these changes inevitably affect brake performance and life. Nowadays, the geometry of brake components tends to be more and more complex, especially in railway applications, with ventilated discs having complex vane/pillar patterns and pads made of small individual friction parts (Fig. 1a). As a result, structural behaviour becomes dependant upon very complex geometries.
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In order to reduce this complexity, a simpler experimental set-up has been developed and the contact interface studies conduced included both, thermal and SEM procedures. A thermal model is developed, which included a volumetric third body (layer) at the interface, with both energy storage and conduction. Finally, comparisons between modelling and experimental results have been conducted. 2. Braking phenomenon 2.1. Macroscopic scale The heat generated in frictional mechanisms, such as brakes and clutches, induces thermal dilatations leading to localised contact areas and hot spots. These localised effects may cause thermal damage and early failures. Anderson and Knapp [8] first proposed a classification of the hot spots observed in automotive braking systems. From experimental investigations on the rubbing surface of brake discs on a testing bench, Dufre´noy et al. [15] gave similar classification for railway brakes, based on thermal infrared measurements (Fig. 2):
Asperity type (Fig. 2-1) result from discrete asperity contacts. Temperatures rise rapidly but briefly, as flash temperatures, on very small areas of the rubbing surface [9]. Rheology of the contact surfaces and the surfaces topography are the main parameters to study this phenomenon [10,11]. Gradients on hot bands (Fig. 2-2) correspond to small contact sites appearing along a single rubbing path. This phenomenon is governed by the contact instabilities and the influence of the pad design, such as thermo elastic instabilities (TEI) [12–14].
Fig. 1. Braking: a multiscale phenomenon ((a) macroscopic [SNCF] and (b) microscopic [10]).
Fig. 2. Classification of hot spots illustrated by thermographs on railway brake discs [15].
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Hot bands (Fig. 2-3) appear as reduced contact friction areas of the pad in the radial direction. They are seen on the disc as narrow rings of high temperatures in the direction of sliding. They can move in the radial direction during braking, according to the evolution of the bearing surfaces. These phenomena, governed by the pad, can be modelled by ‘‘thermo-mechanical loops’’, taking into account the variation of the contact pressure distribution with thermal dilatations and wear of the first bodies [15]. Macroscopic hot spots (MHS) (Fig. 2-4) are large thermal gradients regularly distributed on the disc surface. They are fixed on the disc and appear as a buckling pattern of the disc. This phenomenon reduces drastically the contact surface area with high local temperatures. Recent developments have shown that they can be modelled by successive steps of thermomechanical bending of the disc under specific conditions [16]. MHS correspond to a waviness deformation of the disc exposed to high thermal loading. Regional hot spots (Fig. 2-5) are low thermal gradients on the whole surface of the disc, due to inhomogeneous cooling. Such distributions appear at the end of braking due to thermal diffusion and the non-uniformity of cooling. This phenomenon is mainly governed by the disc.
2.2. Microscopic scale The study of dry friction between rubbing surfaces is generally closely related to the debris rheology of the interface. The debris particles have often been considered as worn particles from the contact area. Numerous experimental studies emphasised the complexity of the mechanisms in sliding contact (third body separation into granular form and sheets of micro-plates, third body flows, etc.) [6,7,17,18]. Many observations showed that the accommodation of relative speed between the first bodies takes place by degradations of the contact surfaces, with cracks and superficial tribological transformations (STT), at a depth ranging from a nanometre to several micrometres. It appears that these STTs are real material flows parallel to the contact surface, resulting from very highly local pressure, temperature and surface contact shear [19]. The third body mechanical behaviour has been widely studied. Most of the published work is based on observations of sliding surfaces after contact opening and understanding of third body elaboration. Few studies (such as [20]) have been devoted to the examination of the interaction between heating areas and third body generation. An obvious difficulty when conducting such experiments is related to the simultaneous observation of the third body formation and the temperature distribution in a sliding contact. In order to address this difficulty in an efficient manner, a simpler experimental set-up has been proposed, which will be explained in detail.
3. The steel–sapphire rings experiment 3.1. Experimental set-up In order to improve the understanding of heat generation mechanisms and surface degradations, an original experimental set-up has been proposed. As shown in Fig. 3, two co-axial rings, made of sapphire and C35 steel are used. The inner and outer radii are 10 and 15 mm, respectively. More details of the set-up are given in [21]. The sapphire ring rotates around its axis, with the steel ring being fixed. The sapphire has been chosen as a material for the rotating ring because of its infrared quasi-transparency (transmission of 80% in the wave length of 3.5–5 mm). In order to obtain an axial heat flow into the rings, two thermal organic fibres insulators are fixed inside and outside the rings (Fig. 3). For temperature measurements, three thermocouples were used at steel ring surface and infrared camera was used to monitor interface temperatures through the sapphire ring. The thermocouples, inserted at the friction surface of the steel ring, are made of four layers: a Kaptons layer bonded on the steel surface, layers of each conductor (copper and Constantans) separated by an electrical insulator. The junction of the two conductors is made at the extremity of the thermocouple (Fig. 3). Three grooves are machined on the inner radius of the steel ring surface at equidistant angular positions. The grooves are curved in order to locate the thermocouple junction at the mean swept radius of the steel ring (Fig. 3). The main difficulty related to the use of an infrared camera, is the non-uniformity and variations of emissivity. A good example of necessary caution when studying thermograms has been given in [21]. Additional measurements with thermocouples have lead to the evaluation of a global emissivity coefficient, which is increasing during the tests as the third body accumulates on the fretting surface. In the general case, infrared measurements show that heating zones are much smaller than the apparent contact surface. 3.2. Results Three types of infrared thermographs may be distinguished according to the shape and the movement of the heating areas. Fig. 4 shows successive infrared thermographs for these three types (the dashed lines correspond to the inner and outer radii of the rings):
Fixed areas (Fig. 4a). The contact is localised on two main areas, with the temperature rise of the order of 200 1C. These areas are fixed from one image to the next. Since the steel ring is not rotating, it can be concluded that these areas are associated with steel contact surface. Moving areas (Fig. 4b), rotating at the same speed as the sapphire ring. There is a 2401 rotation between two successive pictures. Usually, the hot spot areas are more
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Fig. 3. Experimental set-up.
Fig. 4. Infrared measurements.
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elongated than the fixed areas and the temperature increase is lower (of the order of 150 1C). These areas have the same rotational speed as the sapphire ring, which suggest that these areas are associated with the sapphire contact surface. Low speed moving areas (Fig. 4c) with temperatures of the order of 100 1C. During some sequences, angular low speed rotation of the heating zones can be seen, generally after a period of heating (not at the beginning). Fig. 4c illustrates this angular movement of a heating area at three times of a sequence, separated by 29 revolutions between the first and the second and by 56 revolutions between the second and the third. Slight angular movement in the direction of rotation can be seen. Continuous observation of the thermographs shows discontinuous variations of movements for these areas, leading to the conclusion that this phenomenon is most associated with the corresponding third body flows.
Fig. 5 shows macroscopic observations (obtained using SEM) on the steel and sapphire contact surfaces, clearly displaying high surface damage. Numerous traces are observed on the steel contact surface caused by ploughing (Fig. 5a). The contact severity is also visible on the sapphire
contact surface by V oriented cracks centred on the mean contact radius (Fig. 5b). The black arrows indicate the sliding direction (Fig. 5a and b). Fig. 6 shows SEM pictures on the sapphire surface. Two types of the third body can be distinguished. The first one is compact and smooth—the micro-plates in the areas of actual contact. The second one is granular and seems to accumulate in hollows or against material obstacles to form compacted micro-plates. During tests with high loading conditions, heating zones were concentrated, quasi-systematically fixed in rotation. Thermographs show that theses zones appear systematically at the same place and that they are fixed during numerous revolutions. A schematic presentation of the contact area, drawn at the centre of Fig. 7 shows the location of each heating zones through the sapphire during the sequences on the contact surface. Micrographs taken after the tests are shown for each of the depressions, and their position clearly marked. Comparison between these observations and the thermographs (middle of the sequence for thermograph on the top of Fig. 7, and end of sequence for thermograph on the bottom of Fig. 7) show a correlation between the heating zones and depressions observed on the steel surface. Micrographs
Fig. 5. Contact surfaces after the test ((a) C35 steel and (b) sapphire).
Fig. 6. The two-third body types on sapphire ring surface.
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Fig. 7. Comparison between fixed heated areas and depressions on the steel surface.
clearly indicated that depressions have been developed at the steel ring surface systematically on each heating zone. In order to explain the temperature gradient at the interface, and the wear process, two observations will be made:
Non uniformities on the surface—the local areas where the third body accumulates to form compact zones creating contacts. Heat is then generated on these small areas—explaining the high temperatures reached during the tests. Due to thermal expansion, the contact zones
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are taking more contact load and generating more heat, resulting in their distortion. The micro-plates formations can also be interpreted in terms of hot spots [9]. Thermomechanical contact instabilities may appear, depending on the sliding condition—velocity and pressure, leading to nonuniform pressure and local heating. Furthermore, high temperatures reached during the test on compacted zones of the contact surface, suggest the presence of asperity or focal hot spots. Asperity hot spots may be associated with the surface roughness of the materials. Such large temperatures and pressures cause the asperities to wear out rapidly, leaving small marks on the surface. These hot spots are associated with the phenomenon of flash temperatures [22–24]. Focal hot spots are due to thermomechanical instabilities in the sense explained by Barber [12], such as banding hot spots appearing as strips distributed along the sliding surface. They have been clearly observed in the conducted experiments, and lead to more continuous heating than asperity hot spots. The local depressions on the steel surface are most probably the consequence of these focal hot spots. These raised areas are taking the load and therefore most heat is generated here, hence hot spots are being formed in these locations. They are also zones of third body compaction and shearing (such as ‘wear’) is high. Once the experiment (heat generation) stops, these areas will cool down and contract more than the surrounding—that is why ‘depressions’ are created.
These two mechanisms, of third body accumulation and focal hot spots creation, are seen as complementary effects. Focal hot spots are mainly governed by the behaviour of the first bodies (deflections, nonuniform thermal expansion, etc.), leading to third body accumulation. The third body is accumulated and compacted on surface irregularities giving local perturbations leading to thermomechanical instabilities and focal hot spots. This illustrates the necessity to consider interaction between the local thermomechanical behaviour at the interface and the global
interaction between the first bodies (friction pair components). 4. Numerical modelling 4.1. Thermal macroscopic results The thermal problem is generally separately solved by assuming thermal continuity at the interface [25]. In this work, an interface layer, modelled from tribological considerations, is introduced in the thermal problem. The aim is to better understand the mechanisms of heat generation and to give indications of the temperatures near the contact surface. The mechanical aspects are deliberately ignored at this stage. In this model, the third body layer is introduced, between the first bodies, as a continuous thin layer [25]. This layer is volumic with uniform heat generation. Contrary to the classical modelling approach, this model enables energy storage inside the third body (layer) giving thermal distribution across its thickness. The results are mainly sensitive to two parameters: thickness and conductivity of the third body. Taking parameter values from the literature [26], the thickness of 0.01 mm and conductivity of 0.07 W m 1 K 1, results show thermal gradients between the two first body surfaces in agreement with experimental observations found in literature and obtained by the authors [25]. Fig. 8 illustrates these results by showing the temperatures of disc and pad surfaces obtained by the model developed by the authors [25], for an automotive disc brake application. A very high thermal gradient is observed between the two surfaces (of the first bodies) in contact. Pad temperature is higher than disc temperature due to the third body acting as a thermal barrier, creating thermal resistance. As the pad effusivity is lower, its temperature quickly rises in the same form as the heat flux transient evolution. Temperature differences shown in Fig. 8 are very sensitive to the thickness of the third layer (body). Fig. 9
Fig. 8. Disc and pad surface temperature evolution during an automotive braking [25].
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shows the temperature distributions in the third body during braking, for the third body thicknesses of 0.01 mm (Fig. 9a) and 0.1 mm (Fig. 9b). Temperature distribution is represented at different location into the third body. In both cases, a maximal temperature is obtained inside the third body. For thin layers, maximal temperature is obtained near the pad surface because of its lower effusivity. As thickness increases, maximal temperature moves towards the middle of the third body. Results show that introducing a third body leads to higher predicted temperatures. The contact pressure has been assumed to be uniform, but it can be considered that if the areas of real contact are smaller (i.e. assuming localised contact), temperatures would be even higher. This illustrates the necessity to enrich the models by taking into account real contact surface evaluations and material degradations. 4.2. Thermal microscopic results This model has been used in the case of the two rings experimental set-up previously described (Fig. 3). A third body (layer), having a thickness of 5 mm and thermal contact conductance of 7000 W m 2 K 1 has been modelled. The value of thermal conductance is in good
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agreement with those obtained by Day [26]. The contact pressure is assumed to be uniform and thermal analysis conducted in order to predict the temperature gradient at the interface. The numerical results are compared with the experimental thermal measurements obtained in two places: on the top of the third body layer (using the infrared camera—measuring through the sapphire), and on the bottom of the third body layer (using the thermocouples at the steel surface). Fig. 10 shows the comparison of experimental and numerical temperatures. Fig. 10a shows infrared measurements, whilst Fig. 10b shows the evolution of the average temperature obtained by the thermocouples and infrared camera, as well as the average numerical modelling temperature results. For a continuous test, under low loading conditions, the temperature distribution was uniform enough; therefore good agreement with temperature values measured using thermocouples (Fig. 10b) has been achieved. If the experimental measurements are interpreted as temperatures above and under the third body layer, the experimental and numerical thermal gradients are in good agreement in terms of gradients between the two surfaces. However, predicted temperatures are lower than those obtained experimentally. This may be due to the effective contact area assumed in the model as the whole surface,
Fig. 9. Temperature distribution into the third body for two thicknesses ((a) 0.01 mm and (b) 0.1 mm).
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Fig. 10. Comparison between experiments and numerical model ((a) contact surface and (b) thermal results).
Fig. 11. Localised contact ((a) estimated contact surface and (b) comparison between numerical and experimental results).
whereas in reality it seems to be reduced, as shown in the thermograms (Fig. 10a). If contact surface is reduced to half of the apparent one, near the outer radius as shown in Fig. 10a, the numerical results would be closer to the experimental ones. During most of the experiments, the contact area was considerably reduced due to local heating. Analysis of the thermograms allows to estimate the contact surface, which for the test sequence, is ranging from 1/2000th to 1/1000th of the nominal ring friction surface (Fig. 11a). The thermal model with third body has been used in this case, modelling a transient analysis of 10 s duration. The predicted temperatures are compared with the experimentally determined temperatures (thermographs) obtained using infrared camera. Results are presented in Fig. 11b for the two values of estimated contact surface area, Surface 1 assumed to be 1/1000th of the nominal friction surface and Surface 2, equalling 1/2000th of the nominal friction area. The maximal value of temperature on the thermograms during the sequence, marked ‘‘Experiment’’, is also shown. Fig. 11b shows good agreement between the predicted and measured values in terms of transient nominal temperature change during the 10 s event (sequence). Variations of the maximum experimental temperature value are most probably the result of the variation of the
actual contact area during the sequence. The modelling values, chosen to be 1/1000th and 1/2000th of the nominal friction area, can be considered as lower and upper limits of the actual contact area, since maximal experimental temperature values vary between these two temperature curves. 5. Conclusions The developed original experimental set-up, made of two coaxial rings in relative motion, enabled thermal measurements on both sides of the third body. Hot spots have been identified and temperature gradient across the third body accurately measured. Infrared camera and thermocouples have been shown to be an effective tool for this research. Further investigations, using SEM, enabled detailed analysis of friction interfaces of both components, the sapphire and steel rings. Two types of third body layers have been identified, the compact and smooth microplates (where the actual contact occurs) and granular (which seem to accumulate in depressions or against material obstacles). When interface analyses were compared to thermal results, a very important pattern has been discovered. There are clear indications that the hot spots and
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depressions on steel friction surface are directly related. These areas of contact seem to be ‘shrinking’ in height after the application and complete component cooling. The investigation of the third body phenomenon and its influence on interface temperatures (i.e. thermal gradient across the third body layer) has direct relation to the observations made in automotive disc brakes. The thickness of third layer may very, which directly influences thermal gradients. A thermal numerical model, which was also developed, introduced the third body as a uniform layer with energy storage and conduction. The obtained thermal gradients seem to be accurate, when compared with measurements conducted. The results are also similar to those found in literature. In addition, when only a fraction (1/1000th and 1/2000th) of the total friction was considered to be in the actual contact, experimental temperature results were exactly within the predicted range. This indicates that the actual contact area varies during application. Further work will be focused on the introduction of the thermomechanical dimension in the simulations. Contact localisation has to be introduced at two complementary scales: macroscopic with the thermal dilatations of the first bodies (actual components) and microscopic with the rheology of the surfaces and the third body formation and accumulation. References [1] Missori S, Sili A. Optimizing proportions of railway brake discs by temperature transients evaluation. Proc Inst Mech Eng Part F: J Rail Rapid Transit 1988;202:91–9. [2] Newcomb TP. Temperatures reached in disc brakes. J Mech Eng Sci 1960;2(3):167–77. [3] Day AJ, Newcomb TP. The dissipation of frictional energy from interface of an annular disc brake. Proc Inst Mech Eng Part F: J Rail Rapid Transit 1998;198:201–9. [4] Kennedy FE, Ling FF. A thermal, thermoelastic and wear simulation of a high energy sliding contact problem. ASME J Lub Tech 1974;96(2):497–504. [5] Kao T, Richmond JW, Douarre A. Brake disc hot spotting and thermal judder: an experimental and finite element study. Int J Vehicle Des 2000;23:276–96. [6] Berthier Y. Experimental evidence for friction and wear modelling. Wear 1990;115:607–14. [7] Desplanques Y, Degallaix G, Copin R, Berthier Y, A tribometer for the study of materials under braking condition. In: G. Dalmaz, et al., editors. Tribology research: from model experiment to industrial
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