Effects Caliper Design Is Best for a Heavy Duty Use

Disk Brake

The disc brake unit consists of a rotating disc attached to the road-wheel hub and a floating caliper supported on the caliper carrier which is itself bolted to the stub-axle or casing.

From: Advanced Vehicle Technology (Second Edition) , 2002

Brake System Layout Design

Andrew Day , in Braking of Road Vehicles, 2014

Stator Deformation – Disc Brake Caliper and Drum Brake Anchor Plate

Disc brake caliper deflection was briefly discussed in Chapter 5; deformation, displacement and distortion of the caliper arise from caliper deflection under clamp load, caliper twisting under friction drag loading, and thermal deformation. Deformation and deflection of the anchor plate of the type of Simplex drum brake fitted to most passenger cars and light vans are not considered to affect fluid consumption significantly. In all designs of friction brake, stator deformation and deflection depend strongly on the detail design and construction, and the prediction of stator thermal and mechanical deformation by FEA at the detail design stage is essential.

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Clutches and Brakes

Peter R.N. Childs , in Mechanical Design Engineering Handbook, 2014

13.3.1 Disc Brakes

Disc brakes are familiar from automotive applications where they are used extensively for car and motorcycle wheels. These typically consist of a cast iron disc, bolted to the wheel hub. This is sandwiched between two pads actuated by pistons supported in a caliper mounted on the stub shaft (see Figure 13.20). When the brake pedal is pressed, hydraulically pressurized fluid is forced into the cylinders, pushing the opposing pistons and brake pads into frictional contact with the disc. The advantages of this form of braking are steady braking, easy ventilation, balancing thrust loads, and design simplicity. There is no self-energizing action, so the braking action is proportional to the applied force. The use of a discrete pad allows the disc to cool as it rotates, enabling heat transfer between the cooler disc and the hot brake pad. As the pads on either side of the disc are pushed to the disc with equal forces, the net thrust load on the disc cancels.

With reference to Figure 13.21, the torque capacity per pad is given by

(13.24) $T = μ F r_{e}$

where r _e is an effective radius.

The actuating force assuming constant pressure is given by

(13.25) $F = p_{a v} θ \frac{r_{o}^{2} - r_{i}^{2}}{2}$

or assuming uniform wear by

(13.26) $F = p_{\max} θ r_{i} (r_{o} - r_{i})$

where θ (in radians) is the included angle of the pad, r _i is the inner radius of the pad, and r _o is the outer radius of the pad.

The relationship between the average and the maximum pressure for the uniform wear assumption is given by

(13.27) $\frac{p_{a v}}{p_{\max}} = \frac{2 r_{i} / r_{o}}{1 + r_{i} / r_{o}}$

For an annular disc brake, the effective radius is given by Eqn (13.28), assuming constant pressure and Eqn (13.29) assuming uniform wear.

(13.28) $r_{e} = \frac{2 (r_{o}^{3} - r_{i}^{3})}{3 (r_{o}^{2} - r_{i}^{2})}$

(13.29) $r_{e} = \frac{r_{i} + r_{o}}{2}$

For circular pads the effective radius is given by r _e = rδ, where values for δ are given inTable 13.5 as a function of the ratio of the pad radius and the radial location, R/r. The actuating force for circular pads can be calculated using

Table 13.5. Circular pad disk brake design values.

R/r	δ = r _e/r	p _max/p _av
0	1.000	1.000
0.1	0.983	1.093
0.2	0.969	1.212
0.3	0.957	1.367
0.4	0.947	1.578
0.5	0.938	1.875

Source: Fazekas (1972).

(13.30) $F = π R^{2} p_{a v}$

Example 13.6

A caliper brake is required for the front wheels of a sport's car with a braking capacity of 820 N m for each brake. Preliminary design estimates have set the brake geometry as r _i = 100 mm, r _o = 160 mm, and θ = 45°. A pad with a coefficient of friction of 0.35 has been selected. Determine the required actuating force and the average and maximum contact pressures.

Solution

The torque capacity per pad = 820/2 = 410 N m.

The effective radius is $r_{e} = \frac{0.1 + 0.16}{2} = 0.13 m .$

The actuating force is given by $F = \frac{T}{μ r_{e}} = \frac{410}{0.35 \times 0.13} = 9.011 kN .$

The maximum contact pressure is given by

$p_{\max} = \frac{F}{θ r_{i} (r_{o} - r_{i})} = \frac{9.011 \times 10^{3}}{45 \times (2 π / 360) \times 0.1 \times (0.16 - 0.1)} = 1.912 MN / m^{2} .$

The average pressure is given by

$p_{a v} = p_{\max} \frac{2 r_{i} / r_{o}}{1 + r_{i} / r_{o}} = 1.471 MN / m^{2} .$

Example 13.7

A caliper brake is required for the front wheels of a passenger car with a braking capacity of 320 N m for each brake. Preliminary design estimates have set the brake geometry as r _i = 100 mm, r _o = 140 mm, and θ = 40°. Pads with a coefficient of friction of 0.35 have been selected. Each pad is actuated by means of a hydraulic cylinder of nominal diameter 25.4 mm. Determine the required actuating force, the average and the maximum contact pressures, and the required hydraulic pressure for brake actuation.

Solution

The torque capacity per pad is 320/2 = 160 N m

The effective radius is given by $r_{e} = \frac{0.1 + 0.14}{2} = 0.12 m$

The actuation force required is $F = \frac{T}{μ r_{e}} = \frac{160}{0.35 \times 0.12} = 3810 N$

The maximum pressure is

$p_{\max} = \frac{F}{θ r_{i} (r_{o} - r_{i})} = \frac{3810}{40 (2 π / 360) \times 0.1 (0.14 - 0.1)} = 1.364 \times 10^{6} N / m^{2}$

The average pressure is

$p_{a v} = \frac{2 r_{i} / r_{o}}{1 + r_{i} / r_{o}} p_{\max} = \frac{2 \times 0.1 / 0.14}{1 + 0.1 / 0.14} 1.36 \times 10^{6} = 1.137 \times 10^{6} N / m^{2}$

The area of one of the hydraulic cylinders is π0.0127² = 5.067 × 10⁻⁴ m²

The hydraulic pressure required is given by

$p_{hydraulic} = \frac{F}{A_{cylinder}} = \frac{3810}{5.067 \times 10^{- 4}} = 7.519 \times 10^{6} N / m^{2}$

i.e. p _hydraulic ≈ 75 bar.

Full disc brakes, consisting of a complete annular ring pad, are principally used for industrial machinery. The disc clutch equations developed in Section 13.2.1 are applicable to their design. The disc configuration can be designed to function as either a clutch or a brake (a clutch–brake combination) to transmit a load or control its speed.

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Composite materials – modelling, prediction and optimization

Dragan Aleksendrić , Pierpaolo Carlone , in Soft Computing in the Design and Manufacturing of Composite Materials, 2015

Introduction

Disc brakes are safety-critical automotive components that must satisfy tough cost and environmental requirements [231]. According to [232], drivers often evaluate the performance of automotive braking systems in terms of the brake pedal 'feel'. This is one of the first customer touch points during a driving experience, and as such it can be an important contributor to quality perception and customer appeal [233]. The brake pedal feel depends on the synergistic influence of the performance of the braking system and the driver's subjective perception related to the quality of the brake pedal feel. The brake pedal feel gives the driver a perception of the braking dynamics and braking performance of the vehicle. In this context, it is relevant that the brake pedal feel is significantly affected by the properties of the brake friction material, such as, for instance, the compressibility of the friction material.

Motor vehicle brakes are expected to provide high, stable values of the braking torque under different operating conditions. The operating conditions of a brake are determined by the synergistic influence of the brake actuation pressure, the sliding speed and/or the brake interface temperature [234–238]. It is also well known that the interaction in the brake friction pair determines the performance of a brake. Consequently, brake performance is strongly influenced by the behaviour of the friction material. The sensitivity of the performance of a brake friction material to changes in the brake actuation pressure, in synergy with changes in speed and temperature, is particularly important regarding possible corrections of brake performance during a braking cycle. The control and optimization of brake performance and its sensitivity to changes in speed, pressure and temperature during braking is an issue that needs to be analysed better. In other words, the relevant features of the friction material, as obtained after the development and manufacturing phase, could be enhanced further by identification of the influence of the operating conditions on the final performance of the friction pair. Obviously, this could result in remarkable improvement in the performance exhibited. Nowadays, automotive braking systems are expected to offer stable and at the same time maximum performance under all braking conditions, to exactly match the driver's demands. The operation of motor vehicle brakes needs to be intelligently supported in order to optimize their non-linear behaviour. Owing to the strongly non-linear nature of the braking process, the demands imposed on a braking system and especially on the brakes are very complex and cover a wide range of operating conditions [237, 238]. In the braking systems employed on modern motor vehicles, extreme demands are placed on the friction pair and its tribological performance [239–244].

The most important influence on the sensitivity of brake performance to different braking regimes is the interaction in the friction pair. The brake friction material has a significant influence on the contact situation in the friction pair. As a consequence, the properties of friction materials play a crucial role in the driver's perception of braking performance, especially at elevated brake interface temperatures [242–244]. The properties of the brake friction material influence the performance of disc brakes, i.e. their sensitivity to pressure, speed and temperature during a braking cycle, in several different ways. The brake friction material often critically influences the friction level and its variation during braking, the in-stop friction rise, the hot compressibility [245], and so on. On the other hand, drivers expect a relatively constant level of brake performance in various braking regimes, i.e. the highest level of reliability of the brakes. The level of the braking torque in a braking cycle is a very important issue in the use of a vehicle. It should remain at a stable level in a braking cycle and, at the same time, it should be maximized. However, owing to the frequent, high fluctuations in brake performance that occur, especially the rise in the in-stop braking torque, the stability of brake performance is not ensured. This causes a poor brake pedal feel. Accordingly, the driver obtains confusing feedback about the vehicle dynamics and the performance of the braking system. That is why some level of dynamic control and optimization of the brake performance should be provided during a braking cycle. In other words, a functional relationship between the brake pedal travel and the brake performance needs to be established and optimized, in order to match drivers' demands.

Taking into consideration that very complex and highly non-linear phenomena are involved in the braking process, a sufficiently accurate analytical model of brake performance is almost impossible to obtain [245–249]. The reason for that lies in the fact that the coefficient of friction of a brake friction pair has to be considered as a non-linear parameter. This significantly complicates the process of control of the performance of automotive brakes [250]. That is why intelligent methods need to be introduced into the modelling, control and optimization of highly non-linear processes such as the braking process. The implementation of artificial intelligence techniques should allow the development of intelligent systems that incorporate adaptation and learning. An approach to the optimization of the brake performance of a passenger car, based on a synergy of artificial neural networks and genetic algorithms, is proposed and discussed here. The methodologies and outcomes presented here, in the authors' opinion, could contribute to efforts towards the intelligent optimization of the performance of automotive braking systems and to the creation of 'smart braking' abilities. A hybrid neuro-genetic optimization model was developed for this purpose. This dynamic model was used for control and optimization of the braking torque by adjusting the brake actuation pressure according to the previous and current influences of the sliding speed, the pressure and the temperature. A genetic algorithm was coupled to this model to optimize and adjust the actuation pressure of disc brakes to a level which should provide stable and at the same time maximum performance. The optimization process was done according to the brake pedal travel selected by the driver and the inherent capabilities of the brake for the generation of a braking torque under specific braking conditions. This could be considered as a new, intelligent way to develop an advanced brake assistance system for correcting the current brake performance according to its previous behaviour and the selected brake pedal travel.

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Brake system

Heinz Heisler MSc., BSc., F.I.M.I., M.S.O.E., M.I.R.T.E., M.C.I.T., M.I.L.T. , in Advanced Vehicle Technology (Second Edition), 2002

11.2.5 The principle of the disc brake (Fig. 11.4(a, b and c))

The disc brake basically consists of a rotating circular plate disc attached to and rotated by the wheel hub and a bridge member, known as the caliper, which straddles the disc and is mounted on the suspension carrier, stub axle or axle casing (Fig. 11.4(b)). The caliper contains a pair of pistons and friction pads which, when the brakes are applied, clamp the rotating disc, causing it to reduce speed in accordance to the hydraulic pressure behind each piston generated by the pedal effort.

The normal clamping thrust N on each side of the disc (Fig. 11.4(b and c)) acting through the pistons multiplied by the coefficient of friction μ generated between the disc and pad interfaces produces a frictional force F = μN on both sides of the disc. If the resultant frictional force acts through the centre of the friction pad then the mean distance between the centre of pad pressure and the centre of the disc will be

$\frac{R_{2} - R_{1}}{2} = R .$

Accordingly, the frictional braking torque (Fig. 11.4(a)) will be dependent upon twice the frictional force (both sides) and the distance the pad is located from the disc centre of rotation. That is,

$\begin{array}{l} Braking torque = 2 μ N (\frac{R_{2} - R_{1}}{2}) (Nm) \\ i . e . T_{B} = 2 μ N R (Nm) \end{array}$

Example If the distance between the pad's centre of pressure and the centre of disc rotation is 0.12 m and the coefficient of friction between the rubbing faces is 0.35, determine the clamping force required to produce a braking torque of 84 Nm.

$\begin{array}{l} T_{B} = 2 μ N R \\ ∴ Clamping force N = \frac{T_{B}}{2 μ R} \\ = \frac{84}{2 \times 0.35 \times 0.12} \\ = 1000 N \end{array}$

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Applications and case studies

Ian Hutchings , Philip Shipway , in Tribology (Second Edition), 2017

9.3.4 Brakes

Automotive disc brakes decelerate a vehicle by dissipating its kinetic energy in sliding contacts between pads of friction material and brake discs or rotors that rotate with the wheels. An example of a typical brake assembly is shown in Fig. 9.20; sliding between the pads and the discs occurs at about half the speed of the vehicle, and the pads typically cover 10%–15% of the area of the track they describe on the disc.

The coefficient of friction between the pad and disc in a road car is typically about 0.4. Particularly high friction is not important; what is desirable is a stable and consistent value that does not fall significantly as the temperature of the pad is increased (a phenomenon known as brake fade). Large amounts of energy are dissipated at the sliding interface, at up to 30 kW under severe braking conditions. While temperature rises of 150–300 °C are common, significantly higher temperatures occur in both the pads and discs in steep mountain descents or competitive use. In normal passenger cars the discs are usually made from grey cast iron, while the brake pads consist of composites containing many (sometimes 10 or more) different constituents, compacted by hot-pressing into a solid mass with 5%–10% porosity. There is wide variety in the composition of these composite friction materials, but the constituents can be broadly classified as binders, fibres, frictional additives (lubricants and abrasives) and fillers. Table 9.2 provides a summary of the functions of each constituent, as well as some examples.

Table 9.2. Constituents of typical automotive brake friction materials, their functions, and typical examples

Description	Function	Examples
Binder	Form thermally stable matrix	Thermoset phenolic resin
Fibres	Provide mechanical strength	Brass, steel, Kevlar, glass, ceramics (formerly asbestos)
Lubricant	Stabilize friction, especially at high temperature	Graphite, metal sulfides
Abrasive	Increase friction, clean surface film from disc	Zirconia, alumina, chromium oxide, metal silicates
Filler	Improve manufacturability and reduce cost	Mica, vermiculite, barium sulfate

information from Chan, D., Stachowiak, G.W., 2004. Review of automotive brake friction materials. Proc. Inst. Mech. Eng. D J. Automob. Eng. 218, 953–966; Eriksson, M., Bergman, F., Jacobson, S., 2002. On the nature of tribological contact in automotive brakes. Wear 252, 26–36

The nominal pressure between the brake pad and the disc during braking is typically between 1 and 10 MPa. However, the true contact area is only a small proportion of the total pad area, typically 15%–20% for moderate braking loads, distributed over a number of small islands (plateaux) that protrude by a few μm above the level of the rest of the pad surface, as shown in Fig. 9.21. The islands are typically 50–500 μm in size and are formed from protruding hard constituents (such as fibres) surrounded by softer pad materials and wear debris. Isolated contact spots form and disappear rapidly within each island. The frictional behaviour of the pad-disc system thus depends on the composition, microstructure and properties of small areas on the pad as well as the nature and properties of the surface of the disc; both will vary with the history of the contact and the complex response provides a good example of the need for a systems approach in analyzing tribological problems (see Section 8.3).

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Brake Design Analysis

Andrew Day , in Braking of Road Vehicles, 2014

Basic Principles

The modern automotive disc brake is an 'open' type of 'spot' disc brake, i.e. the friction surfaces are not enclosed for protection (a 'dust' shield may be fitted to prevent the ingress of road debris and water spray, but does not enclose the brake). The rotor is a disc that is attached to the wheel hub and rotates with it while the stator, which is attached to the axle or suspension (e.g. the steering knuckle), consists of two opposing brake pads that are held in a 'caliper' and clamped against the disc by the actuation forces. The friction surface of the brake pad only covers a portion (typically no more than 15%) of the rotor friction surface area. As the pads are clamped against the disc by the actuation force, the friction force generated opposes the motion of the disc and slows it down. Because the disc is attached to the road wheel of the vehicle, the vehicle is decelerated. The principle of the disc brake was first patented by Lanchester in 1902 but it was not until the 1950s (Jaguar cars at Le Mans) that their advantages were fully demonstrated. Their resistance to temperature effects, especially fade, compared with even the most advanced drum brakes of the time, enabled racing drivers to brake harder and later than their competitors. This resistance to temperature effects (see Chapter 7) provides greater consistency of performance, which is the main reason why disc brakes have superseded drum brakes for higher duty automotive applications, especially on vehicle front axles.

Disc brake pads have no significant self-servo effect (as explained later), and the friction force generated is usually considered to be directly proportional to the actuation force applied. This means that greater actuation force is required to deliver a particular brake torque output from a disc brake compared with a drum brake, which has a significant self-servo effect. Since the brake pads of a disc brake contact only the side faces of the disc, radial thermal expansion during operation does not affect the torque generated while in the axial direction thermal expansion is small and of little consequence (except in parking brake mode). This permits small running clearances and large area actuators with a high mechanical advantage. Most modern passenger cars and light vans with hydraulically actuated brakes are also fitted with brake boosters to multiply the actuation force that originates from the driver effort (force on the brake pedal). Although the rotors of open disc brakes may be exposed to dirt, dust and water contamination, they have the advantage of being largely self-cleaning via the centrifuge action whereas braking debris tends to accumulate inside a drum brake. Disc rotation also helps dust and gases released during braking to escape. The inherent self-cleaning characteristic can be further assisted by slots in the pads, and grooves or holes in the rotors, although this also reduces their thermal mass (see Chapter 7).

In a hydraulically actuated disc brake (as fitted to a passenger car or light van) slave pistons in the caliper are forced against the pad backplates by hydraulic pressure, generating a normal force at each pad/disc friction interface. The hydraulic seals in the system are designed to provide a small amount of pad retraction via the mechanism of seal 'rollback', so that springs or other devices to move the pads clear of the disc when not being used are not required. This means that threshold pressures, i.e. the hydraulic actuation pressure required to overcome the forces of the springs or retraction mechanisms, are reduced and braking performance becomes more linear. Despite seal 'rollback', disc brake pads do often touch the disc surface while rotating, so residual brake drag losses are not completely eliminated, and some manufacturers are considering positive retraction of the brake pads in their quest for CO₂ emission reduction. For mechanically (or electromechanically) actuated disc brakes, e.g. on larger commercial vehicles that use compressed air for brake actuation, positive pad retraction is required, usually by a spring in the actuator chamber, and therefore threshold pressures are more significant. Threshold pressures and forces should always be included in brake performance calculations. It is possible that hydraulically and/or air-actuated disc brakes will eventually be superseded by electromechanical actuation. Potentially these could provide faster response and better control of running clearances as well as eliminating concerns about the risk of fluid vaporisation and end-of-life recovery of brake fluid in the case of hydraulically actuated systems. Concerns about safety and compliance with legislation coupled with the technological challenges of the high electric current demands have so far been a major inhibitor to the implementation of electromechanical actuation of road vehicle braking systems.

There are two designs of automotive disc brake caliper predominantly in use today. The original design of caliper had a separate actuator for each brake pad either side of the disc in the 'fixed' or 'opposed piston' hydraulic caliper arrangement, illustrated in Figure 5.1.

For each pad there may be one, two or more pistons to ensure that actuation force is uniformly distributed over the pad/disc friction interface. This is especially important for high aspect ratio pads, i.e. where the pad circumferential length is more than about twice its radial width. This design has very few moving parts and the actuation pistons on each side of the rotor give good actuation force equalisation on both sides, but do require brake fluid to be transferred from one side of the caliper to the other, across the 'bridge' part close to the disc, which may be hot. As a result, this design is susceptible to brake fluid vaporisation. The disadvantages include weight and size; the outboard piston requires space inside the road wheel and constrains the position of the disc with respect to the wheel, which is not compatible with the preferred design of many modern front suspensions and steering geometry (Figure 5.2). However, where very wide section tyres permit positioning the brakes deep into large offset wheels, e.g. in high-performance luxury cars, fixed calipers with two or more pistons each side actuating large area high aspect ratio pads to accommodate the power dissipation requirements are popular.

The other, more recent, design of disc brake caliper is the 'sliding' (or 'fist') type caliper (Figure 5.3), which has a fixed carrier attached to the axle casing or mounting, e.g. the steering knuckle, and this carrier is fitted with two rods or pins on which the body of the caliper slides. The caliper body has an actuator on one side only (the inner side) because the force it applies to the inner pad backplate is reacted by the opposing force generated on the pad on the outer side of the disc as the caliper body slides. The applied force and the reaction force are almost equal and opposite; there is a small difference because of friction in the slide pins. Although this is low when new, it can adversely affect force equalisation when the slide pins are worn or damaged, e.g. by corrosion or water ingress and road debris contamination, so effective sealing of the slides is essential. The absence of an outboard actuator allows the rotor to be positioned further outboard, which in turn allows optimum positioning of the suspension lower ball joint to achieve the desired steering geometry. For hydraulic actuation, the brake fluid vaporisation risk is much reduced because there is no fluid path across the caliper.

Mounting the calipers ahead of the centre of the wheel can help caliper (and pad) cooling, but can also deflect cooling air away from the rotor. It can also increase the load carried by the wheel bearings because of the torque reaction developed at the calipers. For these reasons calipers are preferably mounted to the rear of the brake but for the front axles of front wheel drive (FWD) passenger car types of vehicles, packaging requirements of the steering linkage often dictate mounting the calipers forward of the wheel centre.

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Types of friction material formulas

Roberto C. Dante , in Handbook of Friction Materials and their Applications, 2016

3.1 Automotive applications

Automotive brakes, especially disk brakes, are characterized by many denominations following commercial appeal, such as "ceramic materials," non-asbestos organic (NAO), semimetallic, low metal, hybrid and so on. Sometimes some of the characteristics described are really important, such as the lack of asbestos in formulas of NAO materials. However, in most cases, beyond these different names there is no conceptual difference among the several types of formulas. The progressive elimination of asbestos led to a sequence of evolutionary stages. The first challenge that materials engineers had to face was the search for a fiber that could adequately substitute for asbestos. The characteristics that made asbestos so appreciated and widespread in friction materials were its fibrous nature, low density, thermal insulation, mild wear and stable coefficient of friction. At first it seemed complicated to find a substitute for such a multifunctional material since organic fibers—although good insulators in general and in certain cases exhibiting good mechanical strength, such as in the case of aramid fibers—decompose and do not contribute to the stability coefficient of friction. Moreover, the decomposition of these fibers weakens the material structure, causing a wear increment. On the other hand, metallic fibers such as steel fiber, although contributing to mechanical strength, increase the coefficient of friction, have a high density, and are good thermal conductors. All these facts favored the elaboration of complex formulas with many raw materials to replace asbestos. Copper (as powder and fiber) and steel fibers became two important ingredients for friction materials. Although the role of copper is often related to its high thermal conductivity, its contribution to conductivity near the surface is still unclear. However, the transfer of copper to the composite disk tribofilm is a well-known phenomenon. In the United States, The Brake Pad Partnership [ 1] reached consensus that the most effective course of action would be to pursue legislation that reduces the amount of copper used in brakes to an insignificant amount in a phased manner. The resulting bill, SB 346, which became law in September 2010, places a 5% by weight limit on the copper amount used in brakes sold in California by 2021, and reduces that percentage to 0.5% by 2032. The brakes developed to meet these requirements must also meet all applicable safety and performance standards. Other states will follow the example of California legislation.

Other restraints on copper content will likely be introduced in other states in the United States. In any case, the general trend after asbestos banning is to decrease and eliminate all the hazardous and toxic elements or compounds in friction formulas.

The evolution of friction formulas is outlined in Figure 3.1, where the main challenges faced in the last two decades are presented with their solutions. This led to materials with many metals and metal sulfides. The current trend is to reduce the content of heavy metals.

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Braking

James Balkwill , in Performance Vehicle Dynamics, 2018

6.5 Internet-Based Research and Search Suggestions

•: Research the operation of a disc brake. What steps are needed to determine a relationship between brake pedal pressure and braking force developed at the wheel? (You may assume a value for Mu of the disc pad of 0.2 and should balance torques acting on the wheel and consider the ratio of piston sizes between the master cylinder and the calliper. Ignore expansion of the brake lines under pressure.)
•: Power-assisted brakes are often used on road cars that mean that modest brake pedal pressures are able to generate large braking effort. Research the methods used to produce power assistance, explaining how they work. Why are some road cars fitted with electrical systems and others with pneumatic ones driven from the inlet manifold of the engine?
•: Research how the ABS works.
•: Produce a spreadsheet analysis of vehicle braking performance. The approach taken here should be similar in nature to the approach taken to straight-line acceleration in Chapter 4. You should start with a simple approach that determines basic performance. This should then be extended to include more sophisticated methods and considerations. You should be able to develop a similar numerical approach involving the determination of the time taken to slow down through each increment of speed loss. As with the straight-line example, you could include the effect of more subtle effects, such as rotating inertia, wind speed and downforce as a function of speed.

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Friction and Friction Materials

Andrew Day , in Braking of Road Vehicles, 2014

Operational Effects

In principle, Amontons' laws of friction apply for friction materials; however, the coefficient of friction of a resin-bonded composite/cast-iron friction pair does not stay constant and vehicle and brake designers must therefore be prepared to design for variation. It is useful to understand the physical reasons why variation in friction coefficient occurs. The main cause of variation is temperature; as brakes work, they get hot, and the effect of heat is to raise the temperature of the friction material, and very high temperatures can be generated at the friction interface even under relatively low-duty operation because of the low thermal diffusivity of the friction material. The thermoset resin binder thermophysical properties are temperature dependent and those of many of the other constituents will also be changed by temperature. Chemical reactions may occur, and in particular the thermal degradation of the friction material at the interface is known to be an ablation process. The net result is that friction coefficient changes with temperature; typically μ increases slightly up to disc or drum temperatures of about 200–250°C and subsequently decreases as illustrated in Figure 2.1. The precise variation with temperature depends on the friction material.

In brake terms, operating temperature can be usefully defined in terms of the temperature of the brake rotor. There is some debate as to the best way of measuring this; for conventional resin-bonded composite/cast-iron friction pairs rubbing thermocouples can be used, but embedded thermocouples are often preferred, especially for legislative testing, but whichever method is used, consistency is important (see Chapter 9). Friction material manufacturers may prefer to use their own temperature measurement techniques, which are consistent within the company but may not be directly comparable with other methods used elsewhere. More recently, infrared pyrometry has become popular, and provided that the problems of varying surface emissivity can be overcome, this is a good method for identifying surface temperature variations. No method gives an exact measure of the temperatures generated at the actual friction interface, but all can be reliable as a good measure of the temperature generally prevailing for the particular brake operating conditions.

As the brake is applied, the temperature increases and the friction coefficient changes as explained above. To maintain consistency and equivalence in testing, the 'start-of-stop' temperature is usually taken as the reference temperature. Thus, in comparing different applications, the rotor temperature on initial application of the brake is taken as the defining parameter. A typical example of resin-bonded composite friction material performance at different 'start' temperatures, as measured against a cast iron rotor on a small sample friction test rig, is shown in Figure 2.3. These data indicate how the coefficient of friction varies during a test sequence and between test sequences. The test utilised a 10 mm diameter friction material specimen sliding on a cast iron disc rotating at a constant speed equivalent to 7.15 m/s. A constant normal load was applied for 20 s, removed and repeated for 20 applications on a 1-minute cycle. The first application of the 20 was made when the disc had reached the required initial temperature of 80, 100 or 120°C. Natural convection cooling was provided.

The first test (80°C start temperature) indicated μ increasing from about 0.46 to 0.49. The second test (100°C start temperature) showed a fairly stable μ of about 0.48. The third test (120°C start temperature) showed a fairly stable μ reduced to about 0.46. The fourth test returned to a start temperature of 80°C and showed an increase from the 0.46 of the 120°C test to the level indicated in the first 80°C test, but rather surprisingly it then fell back towards the 120°C level. These results show fairly good friction material behaviour for example purposes only; the test was not particularly demanding or long, and the friction pair demonstrated quite high μ.

The reduction of friction coefficient with temperature is commonly referred to as 'fade'. One physical explanation of fade is that volatile organic components from the resin and other constituents generate regions of pressurised vapour or gas at the interface, separating the sliding surfaces and essentially creating pseudo-hydrodynamic sliding conditions. Because such volatile components are in much greater supply in partly cured friction materials, new or 'green' material frictional performance is likely to be noticeably different from that of used friction material, often showing more variation with temperature. For this reason new brake linings should be treated carefully and not exposed to high-duty, high-temperature operation until they have had a chance to bed-in and burnish. In the USA the terms 'burnishing' and 'bedding in' are used interchangeably, with burnishing being the more commonly used. As explained in Chapter 9, bedding-in can be regarded as a process to achieve geometric conformity between the stator and rotor at the friction interface, and burnishing as a process to achieve a steady condition of sliding or tribological contact at the friction interface, which includes exposing new friction material to temperature to fully cure it and release volatiles from the reaction zone (Figure 2.2).

If a friction material is exposed to high-temperature operation sufficient to cause fade, then it would be expected that when the temperature is allowed to return to a lower value, μ will return to its original value as indicated in Figure 2.3. Although this temperature effect is largely reversible, there is often an effect known as 'delayed fade' that can occur and catch out the unwary. In its extreme manifestation, the brakes of a vehicle can be allowed to cool down, but when they are next applied, a low value of μ is generated (see Chapter 9). For resin-bonded composite friction materials paired with a typical cast iron rotor, prolonged sliding at temperatures in excess of around 300°C (depending on the material and the operating conditions) will result in changes in the surface friction material and possibly through the thickness of the pad or lining. The organic components that are there to control the friction and wear characteristics start to degrade thermally, the friction material's performance is significantly affected, and the mechanical strength of the material is reduced. In the extreme, the friction material surface becomes 'denatured' as all the organic constituents are burnt away and only the temperature-resistant components are left (see Figure 2.4). The friction and wear performance deteriorates irreversibly.

Speed can also affect frictional performance. There is a definite transition zone between the static coefficient of friction μ _s and the sliding coefficient of friction. The former is usually higher than the latter, so at very low speeds brakes can overperform, producing vibration effects such as 'creep-groan'. With resin-bonded composite friction materials, speed effects are almost entirely related to temperature distributions and thermal conditions. Higher vehicle speed means higher sliding speed at the friction interface, and a higher rate of energy dissipation. Higher interface temperatures are generated and μ decreases accordingly. This is a phenomenon known as 'speed sensitivity', and is particularly noticeable in heavy commercial vehicles (Day, 1988). The effect of speed and temperature for a typical resin-bonded composite friction material operating against a cast iron on the same small sample test rig as before is shown in Figure 2.5. Note that the speed axis extends from 1000 to 2500 rev/min, and then returns to 1500 rev/min to indicate the repeatability of the frictional performance. It is standard practice to complete a friction material test sequence by repeating a test at the start conditions to check the 'recovery' (see Chapter 9). Data from tests like these can be used to define friction models for use in computational analyses.

There are many other operational and environmental conditions that can influence frictional performance. Water can have two opposing effects: high humidity can raise μ, so that a vehicle's brakes can appear to be very sharp (and noisy) on cold damp mornings, but a few applications can raise the temperature, dry off the water, and bring μ down to the normal operating level. Soaking or immersion in water can reduce frictional performance because of the presence of a lubricating film (liquid or vapour) between the friction surfaces. (It is interesting to note that the controlled introduction of water to a highly thermally loaded friction surface has been used in truck racing to improve brake performance by increasing the heat dissipation through the latent heat of evaporation of the water.)

Most of the μ variation so far considered has been related to high-duty usage. As mentioned above, μ can also be affected by a usage regime of low-duty brake operation, e.g. when the vehicle is driven on short journeys at relatively low speeds with infrequent light braking and resultant low temperatures. This type of usage can result in films being generated on the surface of the friction material and on the mating surface that are associated with low frictional performance (low μ), and is often referred to (in Europe) as 'glazing'. The surface films would need to be removed or replaced before a return to the characteristic steady-state frictional performance can be achieved. The traditional way of dealing with glazing is to apply some high-duty usage, but this does not always work with modern friction materials where the coatings may be particularly tenacious. The term 'glazing' should not be confused with the use of the same term in the USA to describe the result of overheating the friction material, e.g. in high-duty usage or fade and recovery testing.

When a conventional resin-bonded composite disc brake pad or a drum brake lining is newly applied to a cast-iron mating surface (often referred to as 'green' conditions), the tribological conditions at the interface are very different from those 'steady-state' conditions that exist between used and worn brake friction pairs. The process by which steady-state tribological operating conditions are established is termed 'bedding-in' as previously discussed, but it is often called 'burnishing' particularly in the USA, where burnishing is primarily considered to be exposing the friction material to heat cycles to fully cure them and disperse the volatile compounds while bedding-in is a result of the burnishing process. To explain this in more detail, there can be considered to be two aspects of preparing a new brake friction pair for operation:

1.: Through the process of wear, geometric conformity between the two surfaces will be generated so that the whole of the apparent area of the friction surfaces of the stator and rotor are in full contact. This is regarded as 'bedding-in', and if the brake is subjected to heavy-duty usage before bedding-in is complete, thermal damage to the stator and rotor is likely to result because the frictional work is done over a smaller area than either the rotor or stator has been designed for and the work rate or duty level is too high as a result. During this bedding-in process, the friction material (because it has the smaller area of the two components of the friction pair, and is also the less wear-resistant) wears to accommodate the geometric constraints of the brake. Typically a brake lining or pad will not initially make full contact with the brake drum or disc, as evidenced by an unworn region on the rubbing surface, and if this is found on inspection of the friction surfaces, common practice is to estimate the amount of contact and refer to it as 'percentage bedding'. So if an inspection of a disc brake pad indicates that three-quarters of the friction surface is in contact with the disc, this would be recorded as '75% bedded'. Subsequent usage and wear would be expected to bring all the rubbing surfaces into contact to achieve '100% bedded'.
2.: The process of sliding between the friction material and the rotor will cause the friction surfaces to transform by the thermal, mechanical and chemical processes involved in friction until a quasi-steady-state of tribological contact is established at the interface. Transfer films will be generated on the stator and rotor surfaces, which may be polymeric films arising from the binder resin and its components, filler, friction modifiers, etc., or the 'packing' of third body wear debris at the interface, or the modification of surface topography and metallurgy or microstructure. This is regarded as 'burnishing'.

An example of bedding/burnishing is illustrated in Figure 2.6, which shows the friction surface of a passenger car front disc brake pad at three conditions at the start, interim and final stages of the bedding cycle on an inertia dynamometer test (see Chapter 9). It is actually quite difficult to capture the state of bedding in a photograph; the bedded area in the interim condition (the centre photograph of Figure 2.6) is highlighted by light reflection from the shiny contact region, which would be described as burnished. In the 95% bedded state (bottom photograph) the pad friction surface is burnished but is a matt rather than shiny surface, which is more difficult to distinguish. Representative steady-state brake performance is unlikely to be obtained until the rubbing surfaces are both bedded-in and burnished. Studies of contact effects on local heat frictional generation at a brake friction interface, e.g. by Eriksson et al. (2002) and Qi et al. (2004), provide insight into the science of burnishing as well as friction variation in terms of local contact zones, thermal expansion and wear.

As explained earlier, the prediction of the friction and wear characteristics of friction materials from first principles by analysis and calculation is not possible, so development and testing are essential (see Chapter 9). Variations in the μ of disc brake pads and drum brake linings must be anticipated, and good brake and system design can help to minimise the effects of such variations.The value of μ and any associated variation with operating environment or conditions fundamentally defines the 'performance' of the brake, and achieving the required level and consistency of μ forms an essential part of the friction material design and verification. As a general rule, the coefficient of friction μ of a modern friction material can be expected to vary by ±10% from the nominal; thus when a value of μ is used in this book for brake and system design purposes, the performance of the designed system should always be evaluated at these upper and lower limits. As an example, a disc brake pad specified at a μ of 0.4 should be considered to have a friction coefficient of 0.36 ≤ μ ≤ 0.44. Particular usage or environmental conditions may cause the friction material to exhibit performance that might appear to be outside even this ±10% range.

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Out-of-round railway wheels

J. Nielsen , in Wheel–Rail Interface Handbook, 2009

8.4.3 Strategies to reduce wheel roughness

In most cases, a change to disc brakes would significantly reduce rolling noise levels from freight traffic. However, since the number of freight wagons in Europe is about 600 000 and the life of the wagons is up to 30 years, such a replacement would be expensive. Therefore, efforts are made to replace the cast iron brake blocks with blocks of composite or sinter material. Since the physical properties of composite and sinter materials can vary significantly, the design of tailor-made blocks for a specific operation is possible.

One approach is to replace the present cast iron blocks with other types of brake blocks that, in addition, require a modification of the complete braking system. The use of high friction level materials, such as in the type K blocks, would lead to significant noise reductions but, again, this requires large investments by the railway operators. Thus, the introduction of the K blocks is slow, and the focus is instead on the development of new retrofit blocks with low life-cycle costs. The aim is to design new brake blocks, LL blocks, that are completely interchangeable with the existing cast iron brake blocks without changing the braking system. The material cost for K/LL blocks is higher than for cast iron brake blocks, but their life is also longer. ⁴⁹ When composite material blocks are used, a larger part of the generated heat is entered into the wheel than is the case for conventional cast iron brake blocks, and this has implications for the wheel design.

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Effects Caliper Design Is Best for a Heavy Duty Use

Source: https://www.sciencedirect.com/topics/engineering/disk-brake

Effects Caliper Design Is Best for a Heavy Duty Use

Disk Brake

Brake System Layout Design

Stator Deformation – Disc Brake Caliper and Drum Brake Anchor Plate

Clutches and Brakes

13.3.1 Disc Brakes

Composite materials – modelling, prediction and optimization

Introduction

Brake system

11.2.5 The principle of the disc brake (Fig. 11.4(a, b and c))

Applications and case studies

9.3.4 Brakes

Brake Design Analysis

Basic Principles

Types of friction material formulas

3.1 Automotive applications

Braking

6.5 Internet-Based Research and Search Suggestions

Friction and Friction Materials

Operational Effects

Out-of-round railway wheels

8.4.3 Strategies to reduce wheel roughness

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