Advanced Vehicle Technology Episode 3 Part 4 pot

and the road surface. The relationship between the decelerating force and the vertical load on a wheel is known as the adhesion factor ( a ). This is very similar to the coefficient of friction () which occurs when one surface slides over the other, but in the case of a correctly braked wheel, it should always rotate right up to the point of stopping to obtain the greatest retarding resistance. Typical adhesion factors for various road sur- faces are given in Table 11.2. 11.2 Brake shoe and pad fundamentals 11.2.1 Brake shoe self-energization (Fig. 11.2) The drum type brake has two internal semicircular shoes lined with friction material which matches up to the internal rubbing face of the drum. The shoes are mounted on a back plate, sometimes known as a torque plate, between a pivot anchor or wedge type abutment at the lower shoe ends, and at the upper shoe top end by either a cam or hydraulic piston type expander. For simplicity the expander in Fig. 11.2 is represented by two opposing arrows and the shoe linings by two small segmental blocks in the mid-region of the shoes. When the drum is rotating clockwise, and the upper tips of the shoes are pushed apart by the expander force F e , a normal inward reaction force N will be provided by the drum which resists any shoe expansion. As a result of the drum sliding over the shoe linings, a tangential frictional force F t  N will be generated between each pair of rubbing surfaces. The friction force or drag on the right hand shoe (Fig. 11.2) tends to move in the same direction as its shoe tip force F e producing it and accordingly helps to drag the shoe onto the drum, thereby effectively raising the shoe tip force above that of the original expander force. The increase in shoe tip force above that of the input expander force is described as positive servo, and shoes which provide this self-energizing or servo action are known as leading shoes. i:e: F L  F e  F t where F L  leading shoe tip resultant force Likewise considering the left hand shoe (Fig. 11.2) the frictional force or drag F t tends to oppose and cancel out some of the shoe tip force F e produ- cing it. This causes the effective shoe tip force to be less than the expander input force. The resultant reduction in shoe tip force below that of the initial Table 11.2 Adhesion factors for various road surfaces No. Material Condition Adhesion factor 1 Concrete, coarse asphalt dry 0.8 2 Tarmac, gritted bitumen dry 0.6 3 Concrete, coarse asphalt wet 0.5 4 Tarmac wet 0.4 5 Gritted bitumen tarmac wet 0.3 6 Gritted bitumen tarmac greasy 0.25 7 Gritted bitumen, snow compressed greasy 0.2 8 Gritted bitumen, snow compressed dry 0.15 9 Ice wet 0.1 Fig. 11.2 Drum and shoe layout 452 input tip force is described as negative servo, and shoe arrangements which have this de-energizing property are known as trailing shoes. i:e: F T  F e À F t where F T  trailing shoe tip resultant force The magnitude of the self-energizing action is greatly influenced by the rubbing surface tempera- ture, dampness, wetness, coefficient of friction and speed of drum rotation. Changing the direction of rotation of the drum causes the original leading and trailing shoes to reverse their energizing properties, so that the lead- ing and trailing shoes now become trailing and leading shoes respectively. The shoe arrangement shown in Fig. 11.2 is described as a leading-trailing shoe drum brake. Slightly more self-energizing is obtained if the shoe lining is heavily loaded at the outer ends as opposed to heavy mid-shoe loading. 11.2.2 Retarding wheel and brake drum torques (Fig. 11.2) The maximum retarding wheel torque is limited by wheel slip and is given by T w   a WR (Nm) where T w  wheel retarding torque (Nm)  a  adhesion factor W  vertical load on wheel (N) R  wheel rolling radius (m) Likewise the torque produced at this brake drum caused by the frictional force between the lining and drum necessary to bring the wheel to a stand- still is given by T B  Nr (Nm) where T B  brake drum torque (Nm)   coefficient of friction between lining and drum N  radial force between lining and drum (N) r  drum radius (m) Both wheel and drum torques must be equal up to the point of wheel slip but they act in the oppos- ite direction to each other. Therefore they may be equated. i:e: T B  T w Nr   a WR ; Force between lining and drum N   a WR r (N) Example A road wheel has a rolling radius of 0.2 m and supports a load of 5000 N and has an adhesion factor of 0.8 on a particular road surface. If the drum radius is 0.1 m and the coefficient of friction between the lining and drum is 0.4, deter- mine the radial force between the lining and drum. N   a WR r  0:8 Â 5000 Â0:2 0:4 Â 0:1  20 000 N or 20 kN 11.2.3 Shoe and brake factors (Fig. 11.2) If the brake is designed so that a low operating force generates a high braking effort, it is said to have a high self-energizing or servo action. This desirable property is obtained at the expense of stability because any frictional changes dispropor- tionately affect torque output. A brake with little self-energization, while requiring a higher operat- ing force in relation to brake effort, is more stable in operation and is less affected by frictional changes. The multiplication of effort or self-energizing property for each shoe is known as the shoe factor. The shoe factor S is defined as the ratio of the tangential drum drag at the shoe periphery F t to the force applied by the expander at the shoe tip F e . i:e: Shoe factor  Tangential drum force Shoe tip force S  F t F e The combination of different shoe arrangements such as leading and trailing shoes, two leading shoes, two trailing shoes etc. produces a brake factor B which is the sum of the individual shoe factors. Brake factor  Sum of shoe factors i:e: B  (S L  S T ), 2S L ,2S T and (S p  S s ) 11.2.4 Drum shoe arrangements (Fig. 11.3(a±c)) Leading and trailing shoe brakes (Fig. 11.3(a)) If a single cylinder twin piston expander (double acting) is mounted between two shoe tips and the opposite shoe tips react against a fixed abutment, then the leading shoe is forced against the drum in the forward rotation direction, whilst the trailing shoe works against the rotation direction producing 453 Fig. 11.3 (a±d) Various brake shoe arrangements 454 much less frictional drag. Such an arrangement provides a braking effect which is equal in both forward and reverse motion. Rear wheel brakes incorporating some sort of hand brake mechanism are generally of the lead and trailing shoe type. Two leading shoe brakes (Fig. 11.3(b)) By arran- ging a pair of single piston cylinders (single acting) diametrically opposite each other with their pistons pointing in the direction of drum rotation, then when hydraulic pressure is applied, the drum to lining frictional drag force pulls the shoes in the same direction as the shoe tip piston forces, thus causing both shoes to become self-energizing. Such a layout is known as a two leading shoe drum type brake. In reverse, the braking force is reduced due to the drag force opposing the piston tip forces; both shoes in effect then have a trailing action. Two leading shoe brakes are possibly still the most popular light commercial type front wheel brake. Two trailing shoe brakes (Fig. 11.3(c)) If now two separate single acting cylinders are mounted between the upper and lower shoe tips so that both pistons counteract the rotational forward direction of the drum, then the resultant lining drag force will be far less for each shoe, that is, there is a negative servo condition. Brakes with this layout are therefore referred to as two trailing shoe brakes. This arrangement is suitable for application where lining stability is important and a servo assisted booster is able to compensate for the low resultant drag force relative to a given input shoe tip force. A disadvantage of a two trailing shoe brake is for the same brake effect as a two leading shoe brake; much higher hydraulic line pressures have to be applied. Duo-servo shoe brakes (Fig. 11.3(d)) A double acting cylinder expander is bolted to the back plate and the pistons transmit thrust to each adjacent shoe, whereas the opposite shoe tip ends are joined together by a floating adjustment link. On applica- tion of the brake pedal with the vehicle being driven forward, the pistons move both shoes into contact with the revolving drum. The shoe sub- jected to the piston thrust which acts in the same direction as the drum rotation is called the primary shoe and this shoe, when pulled around with the drum, transfers a considerable force to the adjacent shoe tip via the floating adjustment link. This sec- ond shoe is known as the secondary shoe and its initial movement with the drum pushes it hard against the anchor pin, this being permitted by the pistons themselves floating within the cylinder to accommodate any centralization which might become necessary. Under these conditions a com- pounding of both the primary circumferential drag force and that produced by the secondary shoe itself takes place so that a tremendous wedge or self-wrapping effect takes place far in excess of that produced by the two expander pistons alone. These brakes operate equally in the forward or reverse direction. Duo-servo shoe brakes give exception- ally good performance but are very sensitive to changes in shoe lining properties caused by heat and wetness. Because the secondary shoe performs more work and therefore wears quicker than the primary shoe, lining life is equalized as far as possible by fitting a thick secondary shoe and a relatively thin primary shoe. 11.2.5 The principle of the disc brake (Fig. 11.4(a, b and c)) The disc brake basically consists of a rotating cir- cular 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  generated between the disc and pad interfaces pro- duces a frictional force F  N on both sides of the disc. If the resultant frictional force acts through the centre of the friction pad then the mean dis- tance between the centre of pad pressure and the centre of the disc will be R 2 À 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, Braking torque  2N R 2 À R 1 2  (Nm) i:e: T B  2NR (Nm) 455 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. T B  2NR ; Clamping force N  T B 2R  84 2 Â 0:35 Â 0:12  1000 N 11.2.6 Disc brake pad alignment (Fig. 11.4) When the pads are initially applied they are loaded against the disc with uniform pressure, but a small tilt tendency between the leading and trailing pad edges caused by frictional pad drag occurs. In add- ition the rate of wear from the inner to the outer pad edges is not uniform. The bedding-in conditions of the pads will therefore be examined in the two parts as follows: 1 Due to the thickness of the pad there is a small offset between the pad/disc interface and the pad's back plate reaction abutment within the caliper (Fig. 11.4(c)). Consequently, a couple is Fig. 11.4 (a±c) Disc and pad layout 456 produced which tends to tilt the pad into contact with the disc at its leading edge harder compared to the trailing edge. This in effect provides a very small self-energizing servo action, with the result that the wear rate at the leading edge is higher than that at the trailing edge. 2 The circular distance covered by the disc in one revolution as it sweeps across the pad face increases proportionately from the inner to the outer pad's edges (Fig. 11.4(a)). Accordingly the rubbing speed, and therefore the work done, increases from the inner to the outer pad edges, with the result that the pad temperature and wear per unit area rises as the radial distance from the disc centre becomes greater. 11.2.7 Disc brake cooling (Fig. 11.4) The cooling of the brake disc and its pads is achieved mostly by air convection, although some of the heat is conducted away by the wheel hub. The rubbing surface between the rotating disc and the stationary pads is exposed to the vehicle's frontal airstream and directed air circulation in excess of that obtained between the drum and shoe linings. Therefore the disc brake is considerably more stable than the drum brake under continued brake application. The high conformity of the pad and disc and the uniform pressure enable the disc to withstand higher temperatures compared to the drum brake before thermal stress and distortion become pronounced. Because there is far less dis- tortion with discs compared to drums, the disc can operate at higher temperatures. A further feature of the disc is it expands towards the pads, unlike the drum which expands away from the shoe lin- ings. Therefore, when hot, the disc brake reduces its pedal movement whereas the drum brake increases its pedal movement. Cast ventilated discs considerably improve the cooling capacity of the rotating disc (Fig. 11.4(b)). These cast iron discs are in the form of two annular plates ribbed together by radial vanes which also act as heat sinks. Cooling is effected by centrifugal force pushing air through the radial passages formed by the vanes from the inner entrance to the outer exit. The ventilated disc provides consid- erably more exposed surface area, producing some- thing like a 70% increase in convection heat dissipation compared to a solid disc of similar weight. Ventilated discs reduce the friction pad temperature to about two-thirds that of a solid disc under normal operating conditions. Pad life is considerably increased with lower operating tem- peratures, but there is very little effect on the fric- tional properties of the pad material. Ventilated wheels have very little influence on the disc cooling rate at low speeds. At very high speeds a pressure difference is set up between the inside and outside of the wheel which forces air to flow through the vents towards the disc and pads which can amount to a 10% improvement in the disc's cooling rate. The exposure of the disc and pads to water and dirt considerably increases pad wear. The removal of dust shields will increase the cooling rate of the disc and pad assembly but it also exposes the disc and pads to particles of mud, dust and grit which adhere to the disc. This will cause a reduction in the frictional properties of the rubbing pairs. If there has to be a choice of a lower working temperature at the expense of contaminat- ing the disc and pads or a higher working tempera- ture, the priority would normally be in favour of protecting the rubbing surfaces from the atmo- spheric dust and from the road surface spray. 11.2.8 A comparison of shoe factors and shoe stability (Fig. 11.5) A comparison of different brake shoe arrange- ments and the disc brake can be made on a basis of shoe factor, S, or output torque compared against the variation of rubbing coefficient of fric- tions (Fig. 11.5). The coefficient of friction for most linings and pads ranges between 0.35 and 0.45, and it can be seen that within the normal coefficient of friction working range the order of smallest to greatest shoe factor is roughly as follows in Table 11.3. This comparison shows that the torque output (shoe factor) for a single or two trailing shoes is only approximately one-third of the single or two leading shoe brake, and that the combination of a leading and trailing shoe is about twice that of the two trailing shoe, or roughly two-thirds of the two leading shoe arrangement (Fig. 11.5). The disc and pad's performance is very similar to the two trailing shoe layout, but with higher coefficients of friction the disc brake shoe factor rises at a faster rate than that of the two trailing shoe brake. Overall, the duo- servo shoe layout has a superior shoe factor relative to all other arrangements, amounting to roughly five times that of the two trailing shoes and just under twice that of the two leading shoe brake. Conversely, the lining or pad stability, that is, the ability of the shoes or pads to maintain approxi- mately the same shoe factor if there is a small change in the coefficient, due possibly to wetness or an increase in the friction material temperature, alters in the reverse order as shown in Table 11.3. 457 Generally, brakes with very high shoe factors are unstable and produce a relatively large change in shoe factor (output torque) for a small increase or decrease in the coefficient of friction between the rubbing surfaces. Layouts which have low shoe factors tend to produce a consistent output torque for a considerable shift in the coefficient of friction. Because of the instability of shoe layouts with high shoe factors, most vehicle designers opt for the front brakes to be either two leading shoes or disc and pads, and at the rear a leading and trailing shoe system. They then rely on vacuum or hydraulic servo assistance or full power air operation. Thus having, for example, a combined leading and trail- ing shoe brake provides a relatively high leading shoe factor but with only a moderate degree of stability, as opposed to a very stable trailing shoe which produces a very low shoe factor. The proper- ties of each shoe arrangement complement the other to produce an effective and a reliable founda- tion brake. Leadings and trailing shoe brakes are still favoured on the rear wheels since they easily accommodate the hand brake mechanism and pro- duce an extra self-energizing effect when the hand brake is applied, which in the case of the disc and pad brake is not obtainable and therefore requires a considerable greater clamping force for wheel lock condition. 11.2.9 Properties of friction lining and pad materials Friction level (Fig. 11.6) The average coefficient of friction with modern friction materials is between 0.3 and 0.5. The coefficient of friction should be sufficiently high to limit brake pedal effort and to reduce the expander leverage on com- mercial vehicles, but not so high as to produce grab, and in the extreme case cause lock or sprag so that rotation of the drum becomes impossible. The most suitable grade of friction material must be used to match the degree of self-energization created by the shoe and pad configuration and applications. Resistance to heat fade (Fig. 11.6) This is the ability of a lining or pad material to retain its coefficient of friction with an increase in rubbing temperature. The maximum brake torque the lin- ing or pad is to absorb depends on the size and type of brake, gross vehicle weight, axle loading, the front to rear braking ratio and the maximum attainable speed. A good quality material should retain its friction level throughout the working temperature range of the drum and shoes or disc and pads. A reduction in the frictional level in the Fig. 11.5 Relationship of shoe or brake factors and the coefficient of friction for different shoe layouts and the disc brake 458 higher temperature range may be tolerated, pro- vided that it progressively decreases, because a rapid decline in the coefficient of friction could severely reduce the braking power capability when the vehicle is being driven on long descents or subjected to continuous stop-start journey work. The consequences of a fall in the friction level will be greater brake pedal effort with a very poor retardation response. It has been established that changes in the frictional level which occur with rising working temperatures are caused partly by the additional curing of the pad material when it heats up in service and partly because chemical changes take place in the binder resin. Recovery from fade (Fig. 11.6) This is a measure of the ability of a friction material to revert to its original friction level upon cooling after brake lin- ing or pad temperature fade has taken place. The frictional characteristics of a good quality material will return on cooling, even after being subjected to repeatedly severe heating, but an inferior material may have poor recovery and the friction level may be permanently altered. Poor recovery is caused principally by a chemical breakdown in the ingre- dients. This may cause hardening, cracking, flak- ing, charring or even burning of the linings or pads. If the linings or pads are using thermoplastic binder resins a deposit may form on the rubbing surfaces which may distort the friction properties of the material. Resistance to wear (Fig. 11.6) The life of a friction material, be it a lining or pad, will depend to a great extent upon the rubbing speed and pressure. The wear is greatly influenced by the working tempera- ture. At the upper limits of the temperature range, the lining or pad material structure is weakened, so that there is an increase in the shear and tear action at the friction interface resulting in a higher wear rate. Resistance to rubbing speed (Fig. 11.7) The coeffi- cient of friction between two rubbing surfaces should in theory be independent of speed, but it has been found that the intensity of speed does tend to slightly reduce the friction level, particularly at the higher operating temperature range. Poor fric- tion material may show a high friction level at low rubbing speeds, which may cause judder and grab when the vehicle is about to stop, but suffers from a relatively rapid decline in the friction lever as the rubbing speed increases. Resistance to the intensity of pressure (Fig. 11.8) By the laws of friction, the coefficient of friction should not be influenced by the pressure holding the rubbing surfaces together, but with developed friction materials which are generally compounds held together with resin binders, pressure between the rubbing surfaces does reduce the level of fric- tion. It has been found that small pressure increases at relative low pressures produce a marked reduc- tion in the friction level, but as the intensity of Fig. 11.6 Effects of temperature on the coefficient of friction Fig. 11.7 Effects of rubbing speed on the level of friction over the temperature range 459 pressure becomes high the decrease in friction level is much smaller. A pressure-stable lining will pro- duce deceleration proportional to the pedal effort, but pressure-sensitive materials will require a rela- tively greater pedal force for a given braking perfor- mance. Disc brakes tend to operate better when subjected to high rubbing pressures, whereas shoe linings show a deterioration in performance when operating with similar pressures. Resistance to water contamination (Fig. 11.9) All friction materials are affected by water contamina- tion to some extent. Therefore, a safe margin of friction level should be available for wet condi- tions, and good quality friction materials should have the ability to recover their original friction level quickly and progressively (and not behave erratically during the drying out process). A poor quality material may either recover very slowly or may develop over-recovery tendency (the friction level which is initially low due to the wetness rises excessively during the drying out period, falling again as the lining or pad dries out completely). Over-recovery could cause brake-grab and even wheel-lock, under certain driving conditions. Resistance to moisture sensitivity The effects of atmospheric dampness, humidity or dew may increase the friction level for the first few applica- tions, with the result that the brakes may become noisy and develop a tendency to grab for a short time. Moisture-sensitive friction materials should not be used on brakes which have high self- energizing characteristics. Friction materials Materials which may be used for linings or pads generally have their merits and limitations. Sintered metals tend to have a long life but have a relatively low coefficient of friction. Ceramics mixed with metals have much higher coefficient of friction but are very rigid and there- fore must be made in sections. They tend to be very harsh on the drums and disc, causing them to suffer from much higher wear rates than the asbestos- based materials. There has been a tendency to produce friction materials which contain much less asbestos and much more soft metal, such as brass zinc inserts or aluminium granules. Non- asbestos materials are now available which contain DuPont's Kevlar, a high strength aramid fibre. One manufacturer uses this high strength fibre in pulp form as the main body for the friction material, Fig. 11.8 Effects of rubbing pressure on the coefficient of friction Table 11.3 Shoe factor, relative braking power and stability for various brake types Type of brake Shoe factor Relative braking power Stability Single trailing shoe 0.55 Very low Very high Two trailing shoes 1.15 Very low Very high Disc and pad 1.2 Low High Single leading shoe 1.6 High Low Leading and trailing shoes 2.2 Moderate Moderate Two leading shoes 3.0 High Low Duo-servo shoes 5.0 Very high Very low Fig. 11.9 Effects of water contamination on the material's friction recovery over a period of vehicle stops 460 whereas another manufacturer uses a synthetically created body fibre derived from molten blast- furnace slag reinforced with Kevlar for the main body. Some non-asbestos materials do suffer from a drastic reduction in the coefficient of friction when operating in winter temperatures which, if not catered for in the brake design, may not be adequate for overnight parking brake hold. 11.3 Brake shoe expanders and adjusters 11.3.1 Self-adjusting sector and pawl brake shoe mechanism (Fig. 11.10(a, b and c)) With this leading and trailing shoe rear wheel brake layout the two shoes are actuated by opposing twin hydraulic plungers. A downward hanging hand brake lever pivots from the top of the trailing shoe. A toothed sector lever pivots similarly from the top of the leading shoe, but its lower toothed sector end is supported and held in position with a spring loaded toothed pawl. Both shoes are interlinked with a strut bar. Hand brake operation When the hand brake lever is applied the cable pulls the hand lever inwards, causing it to react against the strut. As it tilts it forces the trailing shoe outwards to the drum. At the same time the strut is forced in the opposite direction against the sector lever. This also pushes the leading shoe via the upper pivot and the lower toothed pawl towards the drum. The hand brake shoe expander linkage between the two shoes Fig. 11.10 (a±c) Self-adjusting sector and pawl shoes with forward full hand brake 461 [...]... Mounted on the end of each tappet plunger is a Fig 11.12 (a and b) 11 .3. 4 Wedge shoe adjuster unit (Fig 11. 13) The adjuster housing is made from malleable iron and is spigoted and bolted firmly to the back plate (Fig 11. 13) A hardened steel wedge is employed Strut and cam brake shoe expander Fig 11. 13 Wedge shoe adjuster unit 46 4 with a screw adjuster stem rotating within the wedge which does not rotate,... both the shaft frictional torque and the cam to roller contact torque (Fig 11. 14( c)) 11 .3. 5 S cam shoe expander (Fig 11. 14) Cam expander requirements The object of a cam brake shoe expander (Fig 11. 14( a and b)) is to convert an input camshaft leverage torque into a shoe tip force The shape of the cam profile plays a large part in the effective expansion thrust imposed on the shoe tips as the shoe linings... utilize the full lining thickness which tends to be limited to 19 mm Typical rates of cam lift vary from 0.2 to 0 .4 mm/ deg which correspond to brake factors of about 12 to 16 with the involute cam profile 46 5 Fig 11. 14 (a±c) Air operated foundation brake assembly As the cam lifts (Fig 11. 14( c)) the pressure angle which is made between the cam and roller centre lines and the base circle tangential line... S cam to be generated from an involute spiral (Fig 11. 14( a)) which gives a slight reduction in lift per degree of cam rotation, but maintains a constant cam effective radius so that the shoe tip force always acts in the same direction relative to the cam shoe roller, no matter which part of the cam profile is in contact with the roller (Fig 11. 14( b)) By these means the shoe tip force will remain approximately... the brake shoes move further apart The first part of the outward movement of the leading shoe takes up the clearance between the strut's inner edge and the adjacent side of the rectangular slot formed in the sector lever As the shoe moves further outwards, the strut restrains the sector lever moving with the leading shoe, so that it is forced to swivel Fig 11.11 (a±d) 11 .3. 2 Self-adjusting quadrant and... The partial rotation of the adjusting sleeve unscrews and advances the adjusting screw to a new position This reduces the lining clearance This cycle of events is repeated as the lining wears The self-adjustment action only operates in the forward vehicle direction Once the brake shoes have been installed and manually adjusted no further attention is necessary until the worn linings are replaced 11 .3. 6... 11 .3. 8 Automatic slack adjuster (Fig 11.18(a, b, c and d)) Purpose Once set up automatic slack adjusters need no manual adjustment during the life of the brake linings Self-adjustment takes place whilst the brakes are released (when the clearance between lining and drum exceeds 1. 14 mm) This designed clearance ensures that there is no brake drag and adequate cooling exists for both shoe and drum 46 9... Brake application with half worn linings (Fig 11.11 (b and c)) When the foot brake applied, hydraulic pressure forces the shoe plungers outwards The leading shoe moves out until the clearance between 4 63 raised position and therefore the master cylinder pedal movement will become excessive, providing a warning that the linings need replacing tappet head abutment which guides and supports the twin web... the line will trace out an involute The involute profile may be produced by drawing a base circle and a straight line equal to its circumference and dividing both into the same number of equal parts (Fig 11. 14( a)) From the marked points on the circle draw tangents to represent successive positions of the generated line Step off the unwrapped portion of the Cam design considerations To give the highest... 11.12(b)) the camshaft is rotated, causing the struts to move outwards against the hollow tapper plungers The tappet head abutments force the shoes into contact with the drum, thereby applying the brakes 11 .3. 3 Strut and cam brake shoe expander (Fig 11.12(a and b)) This type of shoe expander is used in conjunction with leading and trailing shoe brakes normally operated by air pressure-controlled brake actuators . whilst the trailing shoe works against the rotation direction producing 4 53 Fig. 11 .3 (a±d) Various brake shoe arrangements 45 4 much less frictional drag. Such an arrangement provides a braking effect. coarse asphalt dry 0.8 2 Tarmac, gritted bitumen dry 0.6 3 Concrete, coarse asphalt wet 0.5 4 Tarmac wet 0 .4 5 Gritted bitumen tarmac wet 0 .3 6 Gritted bitumen tarmac greasy 0.25 7 Gritted bitumen,. required to produce a braking torque of 84 Nm. T B  2NR ; Clamping force N  T B 2R  84 2 Â 0 :35 Â 0:12  1000 N 11.2.6 Disc brake pad alignment (Fig. 11 .4) When the pads are initially applied