TABC Brakes

So, why the change from mechanical braking systems to hydraulic? Even in the mid-1930s, hydraulic brakes were viewed with some suspicion due to poor rubber compounds and brake fluids, barely one generation away from water and glycerine. A failure in just one part of the system could result in total failure with only the mechanical handbrake left.

What were the disadvantages of the mechanical system that pushed the development of hydraulic brakes?

A. The need to balance the braking effort of all the brake drums by mechanical adjustment, both initially and throughout the life of the linings.

B. Complexity in assigning the correct braking effort between the front and rear brakes. *

C. Lost braking effort due to friction in the cable and rod linkages.

D. Wear in pivot points and actuating cams, leading to inconsistent braking amongst the brake drums.

E. The need for essential regular maintenance such as cleaning and lubrication.

F. The cost of producing a robust system that worked reliably. **

* When braking, the weight transfer to the front increases the adhesion of the tyres at the expense of rear tyre adhesion. Typically, the braking effort at the front should be 60% to the rear’s 40%. Balance compensators were available but at increased complexity and cost.

** The high cost of robust mechanical brakes encouraged some manufacturers of cheap cars to opt for hydraulic systems, maybe those who could only afford cheap cars were seen as expendable. Rolls Royce tried a combined system with hydraulic front brakes and mechanically operated rear brakes.

Hydraulic Systems

Main advantage is that the hydraulic pressure is evenly distributed throughout the system and any need for differences in applied force or braking effort can be accomplished by varying the surface area of the pistons in the wheel cylinders. In the case of the TA/B/C, the front wheel cylinder’s piston is 1 inch in diameter, giving a surface area of 0.785 sq. ins. The rear wheel cylinder’s piston is 0.875 inches in diameter, giving a surface area of 0.60 sq. ins. This gives an applied force ratio of 57% to the front and 43% to the rear.

Hydraulic systems are almost frictionless and able to produce increases in mechanical advantage easily. However, in the case of the TA/B/C, the main mechanical advantage is achieved through the brake pedal. The foot pedal section is 10 inches long, whilst that of the master cylinder operating lever is only 2.5 inches, giving a mechanical advantage of 4.

Foot pressure on the pedal can reach 100 lbf under hard braking, producing a force of 400 lbf at the master cylinder piston of cross sectional area 0.6 sq. ins. resulting in a pressure of 670 psi. Such a hydraulic pressure gives a braking force of 525 lbf at the front shoes and 400 lbf at the rear shoes. Pedal pressure might well double in an emergency and I’ve even heard of the brake pedal bending in a crash situation, so a minimum burst pressure rating of 2,000 psi for the brake pipe work is a starting point.

Figure 1 – Brake system schematic.

Modern brake fluids are synthetic and based on Glycol-Ethers, whose main disadvantage is their hygroscopic nature; they readily absorb moisture causing corrosion and forming deposits, which can clog systems, hence the need to change fluids frequently. Expensive mineral based brake fluids were sometimes used on Rolls Royce and Citroen cars.

Brake fluids in wheel cylinders can be exposed to high temperatures from the brake linings, so they need to have a high boiling point temperature to avoid vaporization. The gaseous nature of a vaporized brake fluid is compressible and would cause a serious reduction in applied braking force. Any ingress of moisture through the vented master cylinder under the floor board, would disperse throughout the whole system and reduce the boiling point of the fluid in the wheel cylinders.

Brake fluids are defined by their DOT grade in relation to their boiling temperature for both dry and wet states. Whilst DOT 4 would seem to be better than DOT 3, its rate of boiling point degradation due to moisture content is greater than DOT 3. This would imply that if brake fluids are not frequently changed, one might benefit from use of the lower spec. DOT 3.

DOT (Department of Transport) grading system.

Note 1: The WET specification is for a water based content of about 2%.

Note 2: DOT 5 is a silicone based fluid.

One survey found more than 20% of old cars exceeded a water content of 5% in their brake fluid, which had probably never been changed.

Brake fluid must have a consistent viscosity over a wide temperature range. Corrosion inhibitors are added, but these degrade with time and combined with accumulating moisture levels necessitate frequent changes of non-silicone based fluids.

Water is immiscible with silicone, so any water entering a silicone based system forms pockets, which sink to the bottom. Due to the surface tension of silicone, small water pockets are unlikely to penetrate through any film of silicone that has wetted the base surface of say the Master Cylinder. However, this may not be the case for easily visible pockets of water, which could cause localised corrosion and freezing.

Silicone fluid should only be introduced to a system that is either new or has been flushed out and had new rubber seals and hoses fitted. Due to the fact that it should never be mixed with non-silicone fluids, it is coloured purple.

Liquids are not appreciably compressible, thus brake fluids would need pressures over 20,000 psi to produce a compression of 1 to 2%.

In this respect, Silicone is slightly more compressible, which seems to have justified some complaints about “spongy brakes”. The real issue could be about trapped air bubbles due to poor pouring techniques and a surface tension that resists air bubbles combining, which would have helped when purging the system of air. To avoid introducing air bubbles the following points should be observed:

A. Leave the container of Silicone fluid to stand for a while and do not shake the bottle.

B. Use a funnel, whose end is dipped into any existing fluid in the master cylinder.

C. Pour directly onto the side of the funnel to minimise entrapping any air.

Several attempts at bleeding the brakes may be needed.

Figure 2 – Silicone pouring technique.

Silicone is more compressible than Glycol based fluids particularly at the higher temperatures where Glycol based fluids would have vapour-locked anyway.

Graph 1 – Compressibility of brake fluids.

Only part of the brake system experiences high temperatures whilst the rest of the system remains cool, so the typical compressibility of silicone is less than 1% and that of Glycol based fluid about 0.3%.

The techniques for bleeding brakes can be confusing, the often advised slow pumping of the brake pedal may not sweep air bubbles away from high spots in the system. On the other hand, somewhat surprisingly, vigorous pumping is claimed to induce cavitation, so somewhere between the two speeds seems sensible. Consider using a Gunson Eezibleed kit at about £17 to £18.

Bleeding usually necessitates extra-long strokes of the piston in the master cylinder. The initial strokes should be done cautiously as the piston seals may encounter unpolished sections of the M/C bore.

Brake System Design.

Braking effectiveness is considerably reduced if the wheels are allowed to “lock up” and induce skidding. Maximum braking force is obtained when there is between 10% to 20% slippage between the braked wheel’s circumferential velocity and the road surface. One exception is on loose gravel where a wedge of material builds up in front of the tyre. Avoidance of “locking up” becomes a significant criterion in brake design, as a system powerful enough to be effective at high speeds could result in the wheels “locking up” at modest and low speeds. A proportionate system, where for a given pedal force, the braking effort increases with velocity, is therefore needed.

This is achieved through the self-servo action of the leading brake shoes. The “drag” of the revolving brake drum on the leading edge of the friction lining, causes the leading shoe to be pulled into the drum. The trailing shoe is pushed away, resulting in approximately a 4 to 1 ratio between the braking effects of the shoes.

Figure 3 – Schematic of self-servo action.

If the leading and trailing edges are considered as the tips of levers, rotating about a common focal point, then the frictional force on the leading lever A pulls the lever into the drum, whilst that on B pushes the lever away. This amplification of the braking effort by the self-servo action is not without drawbacks. Any deviation from the assumed coefficient of friction of the linings could result in either “grabbing” of the brakes or a hazardous braking underperformance.

Types of friction lining.

Difficult to find information that isn’t influenced by commercial self-interests; such is the enthusiasm shown by the marketing fraternity that one bit of “puffery” even described their linings as having low wear and anti-friction properties. Original type asbestos based linings were banned in 1998 due to the alarming numbers of garage mechanics suffering from asbestos related diseases, sometimes taking up to 40 years to develop.

Initially, organic fibres such as those from coconuts, bonded by high pressure and resin based adhesives were used. Since then, synthetic linings have been introduced using fibre-glass and Aramid fibres.

Manufacturers tend to keep quiet about their recipe, as up to 20 ingredients may have to be blended to produce an inexpensive high friction material that’s strong, yet flexible enough to be shaped, capable of withstanding high temperatures without fading, non-abrasive to brake drums yet resistant to wear, able to work when wet and not squeal. Some linings include a thin wire mesh of brass, zinc and even steel to stabilise the friction value by conducting heat away from the operating surface and also to strengthen the lining material.

Woven linings with their characteristic hexagonal pattern have been largely augmented by moulded linings, which are easier to produce, although possibly less robust.

There’s no clear indication that brake linings are more suitable for cast iron drums than for steel ones, as the only recommendations that I’ve found suggest that cold-rolled steel harder than Brinnell 180 and close-grained cast iron are equally suitable. However, cast iron is the preferred material due to its greater dimensional stability and hence freedom from warping due to heat; cast aluminium with cast iron linings are also recommended.

Take care when following recommendations on internet Forum sites, for example one recommendation for “Green Gripper Linings” with a higher coefficient of friction u was countered by a warning that such linings wear the drums out.

Dynamic friction coefficient u for most linings is in the range 0.35 to 0.42. Codes on modern linings identify the hot and cold values of u. E is up to 0.35, F is up to 0.45, G is up to 0.55, H is greater than 0.55. These codes do not address the hardness or wear rates.

Soft linings tend to have a higher u, but are more prone to fade at high temperatures, hence being suitable for normal driving. Hard linings suitable for “sporty” driving resist fade, but inherently have a lower u which can be enhanced by adding metal particles. These particles, being abrasive, are bad news for drums no longer being made.

Look out at auto-jumbles for MO/27/1 which could be an asbestos based soft lining. Soft linings need occasional inspection of the front leading shoes as these account for 45% of the braking effect and wear down heavily. Clean drums with a damp rag and then dispose of rag in a plastic bag. Avoid the use an air-line to blow dust out.

There’s a bewildering choice of linings, as some suppliers offer linings intended for industrial uses, such as wind turbine speed brakes. Being able to specify a specific soft lining material, that others have found effective, would need the collective input of many fellow classic car owners whose choice would have been guided by knowledgeable brake lining suppliers. Is this an “action” point?

Figure 4 – Master brake cylinder.

When the piston is pushed forward, fluid is pumped out to operate the wheel cylinders. On the release of the piston, the brake shoe return springs force fluid to return the piston back to its origin. However, if any fluid were to leak out under pressure, the piston would not fully return (the return displacement of the shoes is limited by the adjustment cams) and thus any volume “compensation” of the fluid through the By-Pass port would be prevented. Adding an internal spring pushes the piston right back, creating a partial vacuum that draws fluid past the master cup via the piston’s chamber.

When the piston is at rest against the circlip, the By-Pass port is only just uncovered, allowing the free flow of fluid to and from the reservoir to allow for volume changes due to temperature, which could cause dragging brakes.

Figure 5 – Master cylinder piston.

The larger port enables fluid to enter the piston’s chamber, forming a low pressure fluid seal behind the master cup, with the rear secondary cup sealing against any fluid leakage, whilst the rubber boot guards against any contamination.

However, a problem arises when bleeding the system as any fluid forced out will be drawn back in, air bubbles and all. This is overcome by a dual role check valve.

Figure 6 – Check valve.

When the piston is operated, fluid is forced through the side holes in the check valve cup, displacing the internal rubber sealing cup washer and flowing out through the extension tube. Fluid is then returned by the brake shoe’s return-spring pressure and lifts the cup off the end sealing washer against the pressure of the master cylinder’s internal spring. When the two opposing spring pressures balance, the check valve returns to sit on its sealing washer, leaving a slight back pressure in the system. This helps splay out the sealing cups in the wheel cylinders, reducing leakage and restricting any moisture diffusing through the seals and hoses. It might help to occasionally pump up the brakes during storage.

When bleeding the brakes, the back pressure falls to zero, preventing any fluid passing the spring loaded check valve on the return stroke.

Figure 7 – Push-rod’s knuckle joint.

This may look odd because of the angled adjusting screw, but is well thought out. The crucial issue is the position of the actual ball of the joint. This should be “plumb” or just forward of the pedal spindle, when the knuckle screw is adjusted to give a “free play” of ½ inch at the brake pedal. Limited movements of the ball-end produce a near horizontal thrust vector and displacement of the push-rod. The height of the master-cylinder should be adjusted to be slightly above the ball, as the ball tends to rise up if its displacement is extended due to slack brake shoe adjustment. Check that the push-rod is a suitable length, straight, almost horizontal and not touching the sides of the hollowed out section of the piston over the likely travel of the brake pedal. Some after-market push-rods and master cylinders may deviate from the original dimensions.

Important, check that the ball is secure in its housing. Check brake switch operation, especially after adjusting brakes, as the limited pedal movement may not operate the mechanical switch.

Consider installing a hydraulic switch and LED brake light placed inside the spare wheel. Use a modified banjo bolt on the 3-way adaptor on the end of the master cylinder extension, for the brake switch.

Checking the fluid level in the master cylinder could be made easier if a remote reservoir was located in the engine bay and connected to the screw top of the master cylinder by a tube and elbow adaptor.

Bronze master and wheel cylinders are available with stainless steel pistons, although it’s still possible to sleeve old wheel cylinders.

Handbrake

MOT testers sometimes comment on how effective this is, hardly surprising, given its 10 to 1 mechanical advantage. However, dried-out old grease in the cables can be a problem causing sluggish brake release. In which case, remove the cables and heat up in hot water or with a hot air gun and then pump paraffin through to flush out the old grease. Try 140 grade oil to re-lubricate.

Check the correct assembly of the cable support flanges on the back plate. (see Fig. 8).

Figure 8 – Assembly of cable support flanges on back plate.

Check that the operating arms on the handbrake cross-shaft are nearly vertical when the cables are in tension, otherwise the threaded rods can be bent by being pulled around the cross-shaft. Strangely, there’s no information on adjustment in the “Brown book”.

New rachets and pawls are available from Roger Furneaux and Digby Elliott.

Lining renewal.

Partially drill through the old rivet with a 1/8 inch drill and then punch out with a pin punch. Avoid opening out holes in the brake shoes. After cleaning and degreasing the shoes, inspect for cracks.

Figure 9 – Flaring rivets.

Start riveting from the centre and loosely flare rivets with special “roll-set punch” working outwards. The lining surface can be kept clean by temporarily covering with masking tape. Use a tool clamp to hold down lining and then secure rivets working outwards again to minimise air gaps between the shoe and lining.

Chamfer the leading and trailing edges of the linings. De-grease drums and remove any glaze with emery cloth. Any high spots on the lining can be found by chalking the inside of the drum, and after rotating the drum with a light application of the brakes, the witness marks on the lining can be removed with a rasp type file.

Check points

1. Shoe pivot stud is secure, if loose it can tend to

elongate the hole.

2. Adjustment cams are not loose. (internal spring rusts and breaks; grind away the bit of the shaft that’s been peened over the cam, dis-assemble, fit new spring then clamp cam to the shaft and MIG weld over to retain cam.)

3. Assembly of shoes onto pivot stud. Should be:

Back plate, double coil spring washer (Thackeray type), shoes (offset ends assembled such that the flanges are in line), horse shoe circlip.

Figure 10 – Pivot assembly.

Pipe replacement.

The original pipes were ¼ inch dia. copper. These need to be well supported at regular intervals to avoid the vibration or flexing that causes work hardening, metal fatigue and possible fracture.

Consider using ¼ inch KUNIFER tube, a copper-nickel alloy which resists fatigue and corrosion.

The original pipes had single flared ends. Double flared ends are more effective in sealing, but need two operations using a special flaring tool.

Photo 1 – Flaring tool and bending former.

Figure 11 – Flared ends.

Use a bending former to avoid kinking the tube and remember to put the flare clamping nuts and protective spring over tube before flaring ends. The thread form used on our brakes is 7/16 inch UNF as the system is based on Lockheed. Brass nuts, to avoid corrosion and a special ring spanner are well advised. Pipe lengths and layout are detailed in Doug Pelton’s “From The Frame Up” and stainless steel protective springs are supplied by Roger Furneaux.

Check the pipes are supported by rubber/plastic lined clamps on bare sections of pipe and that the rubber grommets are not worn or perished where the pipes go through the chassis.

Good practice would be to replace synthetic brake fluids every few years and also renew brake hoses. MOT tests on brake rolling roads can show up imbalances due to oil or brake fluid leaks, the calculated efficiency should be in the order of 100% if you give the weight of the car as 787 kg. (Brown book).

I think 100% means the deceleration of the car is numerically equal to the acceleration due to gravity. However, this won’t stop you falling off a cliff!

Vehicle’s total braking force = F. Test weight for vehicle = Gp

Braking efficiency Z = (F / Gp) x 100%

Such an efficiency is not really achievable as the additional weights of fuel, oil, water and the person driving have not been taken into account.

Ed’s Note: Yet another of Eric Worpe’s superb technical articles. The source document for this article was a bundle of flip charts which Eric produced to illustrate his presentation to the T Register’s ‘Rebuild’ seminar held in March. Eric must have spent several hours in writing up his notes and in taking photographs of the flip charts.

As a matter of interest I had occasion to contact Eric recently about the thickness of brake lining material for TA/B/C brakes. I had received an enquiry from a TC owner who was experiencing difficulty in getting his new linings to fit – the drums were getting hot, even after a short run.

My advice to the owner, admittedly based on my Triple-M experience was, 3/16” but I said I would check with Eric. After measuring some different samples, including some original Raybestos linings we came up with a thickness of 4.8mm (3/16” = 4.7625mm). This information was relayed onwards and using this measurement solved the problem.