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Liquid-cooled disc brakes
311
Wear, 34 (1975) 311 - 317 @ Elsevier Sequoia S.A., Lausanne - Printed in the Netherlands
LIQUID-COOLED
DISC BRAKES*
T. P. NEWCOMB and M. EL-SHERBINY Department
of Transport Technology,
University of Loughborough
(Ct. Britain)
(Received May 9, 1975)
Summary The conventional disc brake can reach very high temperatures that seriously impair the brake performance despite the continuing development of better materials used as friction pairs. Lower temperatures can be obtained by using a fluid to remove much of the heat generated by friction. This paper analyses the thermal behaviour of oil-immersed brakes and discusses the factors that influence the friction surface temperature and performance. Finally consideration is given to the use of wet brakes in road vehicles.
Introduction The limiting factor with regard to the performance of brakes is in general, the temperature reached at the sliding surfaces. These high temperatures invariably result in excessive wear of the friction material, loss of performance and, in some cases, structural damage of the surfaces from fracture and thermal fatigue [l] . Experience has shown that a power rating per unit area of friction material or friction surface area can be used as a criterion of failure [ 21. The value of the power rating depends on the type of duty and on the heat dissipation behaviour of the brake. In dry applications the desired rating frequently cannot be obtained owing to space restrictions and this combined with possible inadequate air-cooling results in the brake operating at temperatures above the recommended maximum continuous temperature to give good life from the friction material. Better air-cooling is very difficult to achieve in practice and lower temperatures can more easily be obtained if oil is used to cool the brake. This can conveniently be done in brakes of the multi-disc type as this arrangement gives a more compact brake and greater frictional surface and torque can be obtained by increasing the number of discs. Many friction materials have been developed for wet applications and a brief mention is made of the most suitable for oil-immersed brakes. *Paper presented at the 3rd International Tribology Conference, “Tribology for the Eighties”, Paisley, 22 - 25 September, 1975.
312
In wet applications failure is again related to temperature and, in order to assess the potential of liquid-cooled brakes, an analysis has been made of the heat flow in a multi-disc brake. The results of this work are described in this paper and consideration is given to the use of wet brakes for commercial vehicles. The provision of adequate braking on very heavy commercial vehicles is already a problem and it may well be necessary to use oil-immersed brakes to cope with increased energy dissipations in braking the next generation of commercial vehicles. Materials for oil-immersed
applications
Sintered metal friction materials are best suited to cope with severe duties that involve long braking times. They have exceptional life under normal service conditions and can perform reasonably well under marginal conditions when starved of oil, when temperatures may become very high. The energy dissipation can be up to 2.8 to 3.35W/mm2 before failure for short braking times between 1 - 2 s [ 31. In repeated engagements a power rating of 0.56W/mm2 of friction surface can be tolerated in applications having engagement times up to 20 s [4] and 0.22W/mm2 in single applications lasting up to 40 s [ 51. The p of plain sintered facings is low, about 0.04 - 0.05 but the dynamic 1-1can be increased to about 0.08 by suitably grooving the facing. In use the steel reaction plates should be as flat as possible and should have a smooth finish not greater than 0.76 pm, to give a consistent frictional behaviour. Fluids should be such that oil thickening does not appreciably occur for if a thicker oil film is formed a lower ~1will result. Additives in the oil should be avoided that react with the sintered material to form metal sulphides, etc., as these give rise to a gradual fall-off in p. The wear of the facings may be increased if any wear particles formed are allowed to accumulate in the oil. Any possible clogging of the surfaces if rough could also reduce P. It is therefore essential to provide adequate oil filtration to ensure a consistent performance from the brake. Uneven wear of the facings in an axial direction in a pack of plates also occurs in practice, being greatest nearest the actuating mechanism. This is caused by spline friction and friction from plate separators and results in a fall-off in torque developed by the plates more remote from the operating mechanism and thereby limits the number of plates that can be effectively used in a brake [6]. Optimum performance is obtained with 6 to 8 discs. The oil itself is affected by temperature for as this increases so does the rate of oxidation and decomposition of the oil and because of this it is desirable not to exceed temperatures of 200 “C for prolonged periods. Temperatures can be kept under control by designing to within a certain power rating.
313
Analytical
solution
Temperatures reached in a single brake application To obtain a basic understanding of the thermal behaviour of a brake a model involving heat flow in one dimension will be considered. Heat is generated at the interface of the friction material at the steel reaction member and most flows into the steel member because of its higher thermal conductivity. The fraction z [7] that enters the steel is not constant throughout the application but a representative value can be used without large errors being introduced. The heat flow in the steel member comes from both friction surfaces and at the centre plane of the disc (the origin) can be considered insulated. The surface condition at the semi-thickness d has a heat flux @N-h8 ) where N is a constant rate of working and he is an allowance for cooling from the fluid. In a multi-disc brake the cooling will result from oil passing through the grooves. This boundary condition corresponds in automotive practice to making a brake application to maintain a constant vehicle speed down a long descent and this type of usage is responsible for severe heating of the brakes. Under these conditions the equation to determine the instantaneous surface temperature 13at x = d is given by
(1)
{ 1 +erf($l}] where
K = cr = c = that
thermal conductivity of steel reaction member, t = (a/d2)t’, thermal diffusivity (steel), p = density (steel), specific heat (steel) and t’ = time during braking. This equation shows the temperature build-up depends on a and t for a given z and N. At large values of t eqn. (1) reduces to
e=~f.! [
1+ (m
2n _
n)
(2)
exr-W) 1
and as a is small m - n - 2a and n - -a this simplifies
1
to
1 -exp(.--at) (3) [ If cooling is ignored eqn. (3) reduces to 0 = zNt’/dpc at the end of a brake application [8, 91. In order to illustrate the temperature that can be reached in service a 8 =g
314
N=O-11
W/m’
N=(F055W/mm’
0
2
L
6
8
10
t’lsecondsl
Fig. 1. Variation of temperature rise “C with braking time at three power ratings.
particular example corresponding to one individual application similar to that made while braking on an Alpine descent will be considered. One investigation [lo] has shown that the mean brake application time is approximately 10 s and the highest rate of working on the friction material is fairly close to 0.22W/mm2 (if the friction material is assumed to cover the entire friction surface). If these duties are simulated in a disc brake (d = 1.6 mm) using sintered metal friction material where z = 0.76 [7] (assumed constant) then Fig. 1 shows the temperature time behaviour determined from eqn. (3). In this brake a = 0.006 which is three times the value when the discs operate dry. These results show that about 10% reduction of temperature is obtained at the end of the application compared with that reached when cooling is ignored. A further decrease in temperature of about 10% would occur as z would be lower than 0.76 since the friction material absorbs a greater percentage of the heat generated as the engagement time increases [ 71. The temperature rise at a rate of .working of 0.22W/mm2 would then be around 225 “C and acceptable provided infrequent applications of this severity are made. Some brake applications last longer than 10 s and for continuous operation on a smaller rating it is essential to restrict temperatures to around 200 “C. As the rate of working was less than 0.11W/mm2 most of the time temperatures would not rise above 200 “C for applications less than 20 s duration. If the rate of working is 0.055W/mm2 the temperature rise would be 60 “C after 10 s braking (Fig. 1) and only 170 “C after 40 s braking. Repeated
braking
After cessation of braking the plates cool by the passing of oil over their surfaces and if 0 i is the initial temperature above ambient then the temperature 0 at any instant t can be calculated from the equation
315
-erf(~)}-nexpcnt, (-.?)}I (4)
e = (myni [,,,,,,,,~
{
1+erf
For most values of a the time t is large enough to make the first term in eqn. (4) very small so that
e _= Oi
-2n
(m-n)
exp(nt)
= exp(---at)
(5)
where (m -n) -2a, n m --a and t = crt’/d2 for an actual time t’ before the next brake application. If now the brake makes another application to give another temperature rise 0 i and this braking procedure is continued then after r applications the final temperature 0, above ambient is given by the equation 1 - exp(-rut,)
er =ei
1 - exp(-at,)
i
(6) I the end of one and the start of the next
where t,, is the time between application. The temperature eventually reaches an equilibrium value when the heat loss between applications is equal to the heat produced during an application. This equilibrium temperature 8 ’ is given by
8’ =eO
+
ei
(1 - exp(--at0)
where B. is the ambient temperature. To illustrate the application of these equations, consider repeated applications of the severity discussed previously. If the braking is made at intervals of ten-seconds, making the percentage braking time 50% of the descent time then the final temperature rise would be 48 i at a rate of working of 0.22W/mm2. Against this if applications are made every minute the final temperature rise would be 1.358i. Most applications made at this rate of working and cooling rate would therefore result in temperatures between 1.350 i and 48 i and thus create thermal problems in the brake. If the rate of working is reduced to 0.055W/mm2 then a considerable reduction in temperature to around 250 “C w’ould occur at continuous operation of the brake. Application
to commercial
vehicles
As a failure of a friction brake is primarily thermal in origin consideration is now given to the heat dissipation in commercial vehicle brakes. The temperatures involved in braking can be divided into the transient temperature, that is the rapidly varying temperature during the course of a
316
brake application, and the bulk or soak temperature which will increase slowly during a series of stops and creates a more serious thermal problem. The transient temperature will depend on, among other things, the area of friction surface and therefore the latter should be as large as possible to reduce the rate of working. On current commercial vehicles the area of drum friction surface is normally between 54 - 57 mm2 per kg weight of vehicle and although this is less than in a car brake thermal problems do not normally occur in individual brake applications made from high speeds. The bulk drum temperature build-up on a descent depends on the rate of working, frequency of braking, thermal capacity and cooling rate. If braking continues for long periods the final temperature is very dependent on the cooling rate. Commercial vehicle drums have a low cooling rate that makes it advisable not to exceed a continuous rate of energy dissipation of approximately 15kW per drum to avoid excessive temperature and loss of performance. As an example, if a vehicle of weight 16288 kg fitted with four brakes is continuously braked to maintain a steady speed of 30 km down a 1 in 16.7 gradient 82.0kW the steady power and drum temperatures can reach over 400 “C for long periods. Lower temperatures can be attained by using a fluid to carry more heat away from the friction surfaces to be conveniently dissipated eventually to the atmosphere elsewhere in the braking system. In a multi-disc brake if one considers a six-plate brake, each plate having two annuli friction surfaces of o. d. 385 mm and i. d. 330 mm then the total friction surface area from four brakes is 1484 X lo3 mm2. If these brakes are fitted to a vehicle of weight 16288 kg being braked from 96 km/h at a deceleration of 0.6 g the average power dissipated is 1283 kW which is equivalent to a power rating of 0.86 W/mm2. This rating is well within the capabilities of a sintered metal friction material and as used in liquid-cooled brakes they would have little difficulty in coping with individual brake applications involving high energy dissipations. The brake also has adequate thermal capacity to withstand the heat generated in the event of failure of the oil-cooling system. On the long descent mentioned previously the power rating would be 0.055 W/mm2 and when continuously working at this rate the final temperature rise would be 250 “C. This temperature is much less than when the brake was dry and the temperature could be made lower still by increasing the area of friction surface or by better cooling. Better cooling will also allow higher power ratings to be used. Liquid-cooled brakes therefore offer considerable potential for development for commercial vehicle applications. These brakes can have greater energy capacities than dry brakes and are flexible enough to vary the number of plates to cover a wide range of duties without loss of performance. They have the further advantages of not requiring adjustment for lining wear and their life can be made such that replacement is not necessary before an engine overhaul. Oil-immersed brakes will therefore extend the energy range that can be catered for by a friction brake with ease of serviceability and enable them to meet the ever-increasing demands made by increased vehicle loads and speeds.
317
References 1 M. J. Neale, Failure of friction surfaces, Tribology Handbook, Butterworths, London, 1973, Section E5. 2 T. P. Newcomb and R. T. Spurr, Temperature as a criterion of failure in brakes and clutches, First European Tribology Congress, Instn. Mech. Engrs., 1973, Paper C266/73, p. 71. 3 T. P. Newcomb and R. T. Spurr, The interaction between friction materials and lubricants, Wear, 24 (1973) 69. 4 A. Jenkins, T. P. Newcomb and R. C. Parker, Friction facings for clutches, Conference on Drive Line Engineering, Jersey, Proc. Instn. Mech. Engrs., 184 (31) (1969 - 70) 39. 5 D. C. Quick and F. Sippel, Design and development of wet disc brakes for agricultural tractors, SAE Trans. (1973) U.S.A. Paper number 730863. 6 T. P. Newcomb and H. E. Merritt, Effect of spline friction on the torque capacity and interface temperatures reached during a multi-disc engagement, J. Mech. Eng. Sci., 4 (4) (1962) 353. 7 T. P. Newcomb, Interfacial temperatures and the distribution of heat between bodies in sliding contact, Int. Heat Transfer Conf., ASME, 1961, p. 77. 8 T. P. Newcomb and R. T. Spurr, Braking of Road Vehicles, Chapman and Hall, London, 1967, p. 140. 9 Ting-Long Ho, M. B. Peterson and F. F. Ling, Effect of frictional heating on brake materials, Wear, 30 (1974) 73. 10 A. D. M. Frood, D. K. Mackenzie and T. P. Newcomb, Brake usage in a heavy vehicle, Proc. Auto. Div. Instn. Mech. Engrs., 1960 - 1, p. 351.
Wear, 34 (1975) 311 - 317 @ Elsevier Sequoia S.A., Lausanne - Printed in the Netherlands
LIQUID-COOLED
DISC BRAKES*
T. P. NEWCOMB and M. EL-SHERBINY Department
of Transport Technology,
University of Loughborough
(Ct. Britain)
(Received May 9, 1975)
Summary The conventional disc brake can reach very high temperatures that seriously impair the brake performance despite the continuing development of better materials used as friction pairs. Lower temperatures can be obtained by using a fluid to remove much of the heat generated by friction. This paper analyses the thermal behaviour of oil-immersed brakes and discusses the factors that influence the friction surface temperature and performance. Finally consideration is given to the use of wet brakes in road vehicles.
Introduction The limiting factor with regard to the performance of brakes is in general, the temperature reached at the sliding surfaces. These high temperatures invariably result in excessive wear of the friction material, loss of performance and, in some cases, structural damage of the surfaces from fracture and thermal fatigue [l] . Experience has shown that a power rating per unit area of friction material or friction surface area can be used as a criterion of failure [ 21. The value of the power rating depends on the type of duty and on the heat dissipation behaviour of the brake. In dry applications the desired rating frequently cannot be obtained owing to space restrictions and this combined with possible inadequate air-cooling results in the brake operating at temperatures above the recommended maximum continuous temperature to give good life from the friction material. Better air-cooling is very difficult to achieve in practice and lower temperatures can more easily be obtained if oil is used to cool the brake. This can conveniently be done in brakes of the multi-disc type as this arrangement gives a more compact brake and greater frictional surface and torque can be obtained by increasing the number of discs. Many friction materials have been developed for wet applications and a brief mention is made of the most suitable for oil-immersed brakes. *Paper presented at the 3rd International Tribology Conference, “Tribology for the Eighties”, Paisley, 22 - 25 September, 1975.
312
In wet applications failure is again related to temperature and, in order to assess the potential of liquid-cooled brakes, an analysis has been made of the heat flow in a multi-disc brake. The results of this work are described in this paper and consideration is given to the use of wet brakes for commercial vehicles. The provision of adequate braking on very heavy commercial vehicles is already a problem and it may well be necessary to use oil-immersed brakes to cope with increased energy dissipations in braking the next generation of commercial vehicles. Materials for oil-immersed
applications
Sintered metal friction materials are best suited to cope with severe duties that involve long braking times. They have exceptional life under normal service conditions and can perform reasonably well under marginal conditions when starved of oil, when temperatures may become very high. The energy dissipation can be up to 2.8 to 3.35W/mm2 before failure for short braking times between 1 - 2 s [ 31. In repeated engagements a power rating of 0.56W/mm2 of friction surface can be tolerated in applications having engagement times up to 20 s [4] and 0.22W/mm2 in single applications lasting up to 40 s [ 51. The p of plain sintered facings is low, about 0.04 - 0.05 but the dynamic 1-1can be increased to about 0.08 by suitably grooving the facing. In use the steel reaction plates should be as flat as possible and should have a smooth finish not greater than 0.76 pm, to give a consistent frictional behaviour. Fluids should be such that oil thickening does not appreciably occur for if a thicker oil film is formed a lower ~1will result. Additives in the oil should be avoided that react with the sintered material to form metal sulphides, etc., as these give rise to a gradual fall-off in p. The wear of the facings may be increased if any wear particles formed are allowed to accumulate in the oil. Any possible clogging of the surfaces if rough could also reduce P. It is therefore essential to provide adequate oil filtration to ensure a consistent performance from the brake. Uneven wear of the facings in an axial direction in a pack of plates also occurs in practice, being greatest nearest the actuating mechanism. This is caused by spline friction and friction from plate separators and results in a fall-off in torque developed by the plates more remote from the operating mechanism and thereby limits the number of plates that can be effectively used in a brake [6]. Optimum performance is obtained with 6 to 8 discs. The oil itself is affected by temperature for as this increases so does the rate of oxidation and decomposition of the oil and because of this it is desirable not to exceed temperatures of 200 “C for prolonged periods. Temperatures can be kept under control by designing to within a certain power rating.
313
Analytical
solution
Temperatures reached in a single brake application To obtain a basic understanding of the thermal behaviour of a brake a model involving heat flow in one dimension will be considered. Heat is generated at the interface of the friction material at the steel reaction member and most flows into the steel member because of its higher thermal conductivity. The fraction z [7] that enters the steel is not constant throughout the application but a representative value can be used without large errors being introduced. The heat flow in the steel member comes from both friction surfaces and at the centre plane of the disc (the origin) can be considered insulated. The surface condition at the semi-thickness d has a heat flux @N-h8 ) where N is a constant rate of working and he is an allowance for cooling from the fluid. In a multi-disc brake the cooling will result from oil passing through the grooves. This boundary condition corresponds in automotive practice to making a brake application to maintain a constant vehicle speed down a long descent and this type of usage is responsible for severe heating of the brakes. Under these conditions the equation to determine the instantaneous surface temperature 13at x = d is given by
(1)
{ 1 +erf($l}] where
K = cr = c = that
thermal conductivity of steel reaction member, t = (a/d2)t’, thermal diffusivity (steel), p = density (steel), specific heat (steel) and t’ = time during braking. This equation shows the temperature build-up depends on a and t for a given z and N. At large values of t eqn. (1) reduces to
e=~f.! [
1+ (m
2n _
n)
(2)
exr-W) 1
and as a is small m - n - 2a and n - -a this simplifies
1
to
1 -exp(.--at) (3) [ If cooling is ignored eqn. (3) reduces to 0 = zNt’/dpc at the end of a brake application [8, 91. In order to illustrate the temperature that can be reached in service a 8 =g
314
N=O-11
W/m’
N=(F055W/mm’
0
2
L
6
8
10
t’lsecondsl
Fig. 1. Variation of temperature rise “C with braking time at three power ratings.
particular example corresponding to one individual application similar to that made while braking on an Alpine descent will be considered. One investigation [lo] has shown that the mean brake application time is approximately 10 s and the highest rate of working on the friction material is fairly close to 0.22W/mm2 (if the friction material is assumed to cover the entire friction surface). If these duties are simulated in a disc brake (d = 1.6 mm) using sintered metal friction material where z = 0.76 [7] (assumed constant) then Fig. 1 shows the temperature time behaviour determined from eqn. (3). In this brake a = 0.006 which is three times the value when the discs operate dry. These results show that about 10% reduction of temperature is obtained at the end of the application compared with that reached when cooling is ignored. A further decrease in temperature of about 10% would occur as z would be lower than 0.76 since the friction material absorbs a greater percentage of the heat generated as the engagement time increases [ 71. The temperature rise at a rate of .working of 0.22W/mm2 would then be around 225 “C and acceptable provided infrequent applications of this severity are made. Some brake applications last longer than 10 s and for continuous operation on a smaller rating it is essential to restrict temperatures to around 200 “C. As the rate of working was less than 0.11W/mm2 most of the time temperatures would not rise above 200 “C for applications less than 20 s duration. If the rate of working is 0.055W/mm2 the temperature rise would be 60 “C after 10 s braking (Fig. 1) and only 170 “C after 40 s braking. Repeated
braking
After cessation of braking the plates cool by the passing of oil over their surfaces and if 0 i is the initial temperature above ambient then the temperature 0 at any instant t can be calculated from the equation
315
-erf(~)}-nexpcnt, (-.?)}I (4)
e = (myni [,,,,,,,,~
{
1+erf
For most values of a the time t is large enough to make the first term in eqn. (4) very small so that
e _= Oi
-2n
(m-n)
exp(nt)
= exp(---at)
(5)
where (m -n) -2a, n m --a and t = crt’/d2 for an actual time t’ before the next brake application. If now the brake makes another application to give another temperature rise 0 i and this braking procedure is continued then after r applications the final temperature 0, above ambient is given by the equation 1 - exp(-rut,)
er =ei
1 - exp(-at,)
i
(6) I the end of one and the start of the next
where t,, is the time between application. The temperature eventually reaches an equilibrium value when the heat loss between applications is equal to the heat produced during an application. This equilibrium temperature 8 ’ is given by
8’ =eO
+
ei
(1 - exp(--at0)
where B. is the ambient temperature. To illustrate the application of these equations, consider repeated applications of the severity discussed previously. If the braking is made at intervals of ten-seconds, making the percentage braking time 50% of the descent time then the final temperature rise would be 48 i at a rate of working of 0.22W/mm2. Against this if applications are made every minute the final temperature rise would be 1.358i. Most applications made at this rate of working and cooling rate would therefore result in temperatures between 1.350 i and 48 i and thus create thermal problems in the brake. If the rate of working is reduced to 0.055W/mm2 then a considerable reduction in temperature to around 250 “C w’ould occur at continuous operation of the brake. Application
to commercial
vehicles
As a failure of a friction brake is primarily thermal in origin consideration is now given to the heat dissipation in commercial vehicle brakes. The temperatures involved in braking can be divided into the transient temperature, that is the rapidly varying temperature during the course of a
316
brake application, and the bulk or soak temperature which will increase slowly during a series of stops and creates a more serious thermal problem. The transient temperature will depend on, among other things, the area of friction surface and therefore the latter should be as large as possible to reduce the rate of working. On current commercial vehicles the area of drum friction surface is normally between 54 - 57 mm2 per kg weight of vehicle and although this is less than in a car brake thermal problems do not normally occur in individual brake applications made from high speeds. The bulk drum temperature build-up on a descent depends on the rate of working, frequency of braking, thermal capacity and cooling rate. If braking continues for long periods the final temperature is very dependent on the cooling rate. Commercial vehicle drums have a low cooling rate that makes it advisable not to exceed a continuous rate of energy dissipation of approximately 15kW per drum to avoid excessive temperature and loss of performance. As an example, if a vehicle of weight 16288 kg fitted with four brakes is continuously braked to maintain a steady speed of 30 km down a 1 in 16.7 gradient 82.0kW the steady power and drum temperatures can reach over 400 “C for long periods. Lower temperatures can be attained by using a fluid to carry more heat away from the friction surfaces to be conveniently dissipated eventually to the atmosphere elsewhere in the braking system. In a multi-disc brake if one considers a six-plate brake, each plate having two annuli friction surfaces of o. d. 385 mm and i. d. 330 mm then the total friction surface area from four brakes is 1484 X lo3 mm2. If these brakes are fitted to a vehicle of weight 16288 kg being braked from 96 km/h at a deceleration of 0.6 g the average power dissipated is 1283 kW which is equivalent to a power rating of 0.86 W/mm2. This rating is well within the capabilities of a sintered metal friction material and as used in liquid-cooled brakes they would have little difficulty in coping with individual brake applications involving high energy dissipations. The brake also has adequate thermal capacity to withstand the heat generated in the event of failure of the oil-cooling system. On the long descent mentioned previously the power rating would be 0.055 W/mm2 and when continuously working at this rate the final temperature rise would be 250 “C. This temperature is much less than when the brake was dry and the temperature could be made lower still by increasing the area of friction surface or by better cooling. Better cooling will also allow higher power ratings to be used. Liquid-cooled brakes therefore offer considerable potential for development for commercial vehicle applications. These brakes can have greater energy capacities than dry brakes and are flexible enough to vary the number of plates to cover a wide range of duties without loss of performance. They have the further advantages of not requiring adjustment for lining wear and their life can be made such that replacement is not necessary before an engine overhaul. Oil-immersed brakes will therefore extend the energy range that can be catered for by a friction brake with ease of serviceability and enable them to meet the ever-increasing demands made by increased vehicle loads and speeds.
317
References 1 M. J. Neale, Failure of friction surfaces, Tribology Handbook, Butterworths, London, 1973, Section E5. 2 T. P. Newcomb and R. T. Spurr, Temperature as a criterion of failure in brakes and clutches, First European Tribology Congress, Instn. Mech. Engrs., 1973, Paper C266/73, p. 71. 3 T. P. Newcomb and R. T. Spurr, The interaction between friction materials and lubricants, Wear, 24 (1973) 69. 4 A. Jenkins, T. P. Newcomb and R. C. Parker, Friction facings for clutches, Conference on Drive Line Engineering, Jersey, Proc. Instn. Mech. Engrs., 184 (31) (1969 - 70) 39. 5 D. C. Quick and F. Sippel, Design and development of wet disc brakes for agricultural tractors, SAE Trans. (1973) U.S.A. Paper number 730863. 6 T. P. Newcomb and H. E. Merritt, Effect of spline friction on the torque capacity and interface temperatures reached during a multi-disc engagement, J. Mech. Eng. Sci., 4 (4) (1962) 353. 7 T. P. Newcomb, Interfacial temperatures and the distribution of heat between bodies in sliding contact, Int. Heat Transfer Conf., ASME, 1961, p. 77. 8 T. P. Newcomb and R. T. Spurr, Braking of Road Vehicles, Chapman and Hall, London, 1967, p. 140. 9 Ting-Long Ho, M. B. Peterson and F. F. Ling, Effect of frictional heating on brake materials, Wear, 30 (1974) 73. 10 A. D. M. Frood, D. K. Mackenzie and T. P. Newcomb, Brake usage in a heavy vehicle, Proc. Auto. Div. Instn. Mech. Engrs., 1960 - 1, p. 351.