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Technical matters 1927-32

Diesel engine background

The diesel engine takes its name from its inventor Dr Rudolph Diesel, 1858-1913, a German engineer and man of science. He took out his original patents in 1892, the first successful diesel engine was built in around 1895 and the diesel went on to become a commercial success. In developing the diesel engine into a working device, Dr Diesel had to depart considerably from some of the features in the original patent specification as these were found to be unworkable. However, this did not prevent the diesel engine from being developed into a highly satisfactory prime mover.

The main difference between a diesel engine and a petrol engine lies in the way in which the fuel is ignited for the power stroke. A petrol engine uses an electric spark ignition system, whereas the diesel engine uses compression ignition. Generations of cyclists have discovered that a bicycle pump becomes warm when they inflate their tyres, due to the fact that when air is compressed its temperature rises, and this feature is used as the means of ignition in a diesel engine. To this end, a diesel engine has a high enough compression ratio to raise the air temperature in the cylinder head to a level able to ignite injected fuel. For instance, a compression ratio of 14.2 to 1 will result in an air pressure in the cylinder head of around 480psi (33.1bar) gauge, and a corresponding temperature of 900 to 950 degrees F (482 to 510 degrees C). Thermal efficiency, i.e. fuel efficiency, increases with rising compression ratio, therefore the diesel engine’s high CR compared to that of other IC piston engines (e.g. petrol) gives the diesel greater fuel economy. The difference in diesel and petrol-engine CRs is illustrated by Thornycroft practice in 1932, the firm’s diesel engines had CRs of around 14.5 to 1 whereas CRs of its petrol engines were only about 5 to 1.

Discounting miniature two-stroke diesel engines which were once widely used for powering model aircraft, the cycle of operations in diesel engines normally includes the three following features:

  1. Intake and compression of pure air only;
  2. Sufficient compression to vaporise and ignite the fuel as it is admitted to the cylinder head, by the heat of compression;
  3. Fuel injection timed so that combustion starts at the point of maximum compression, and continues at such a rate as to maintain roughly constant pressure during the first part of the working stroke. Injection and combustion cease for the remainder of the working stroke, during which the burnt gas in the cylinder works expansively (with falling pressure) until the exhaust valve/port opens.

The constant pressure combustion referred to in item 3 departs from Dr Diesel’s original patent specification, which proposes that combustion should take place at constant temperature. His original proposal for constant temperature combustion suggests that he intended the diesel engine to operate, as far as possible, on the Carnot heat engine cycle. This cycle was conceived by Frenchman Sadi Carnot, who wrote a paper on heat engines in 1824, and his cycle has the highest thermal efficiency of the various heat engine cycles available. However, Dr Diesel abandoned his original patent proposal and adopted (as near as possible) constant pressure combustion instead, which formed the basis of what became known as the Diesel cycle.

Crankshaft torsional vibrations

Twisting forces acting on a crankshaft comprise 1) gas forces, due to gas pressures within the cylinders, and 2) inertial forces, due to the rapid reciprocating movements of the pistons, connecting rods and gudgeon pins. Since these forces vary over each revolution, then so will the amount of crankshaft twist, i.e. the crankshaft is subject to torsional vibrations due to gas and inertia forces. If the frequency of a vibration forced upon the crankshaft matches the crankshaft’s natural torsional frequency, then, in extreme cases, torsional displacements will rapidly build up, leading to breakage. Torsional displacement is highest at the free end of the crankshaft (i.e. the non-flywheel end) due to the flywheel’s mass.

Because the connecting rod forms a varying angle with the cylinder axis during the rotation of the crankshaft, the inertia force due to each piston/conrod/gudgeon pin assembly comprises a series of concurrent time-varying forces, known as harmonics, each of which has a different frequency and maximum value (amplitude). Amplitude diminishes as frequency rises and so higher harmonics are sometimes ignored when calculating inertia forces.

When six-cylinder engines were introduced, crankshaft torsional vibration problems arose because of 1) the frequencies of the twisting forces, due to the angular disposition of the crank throws and number of cylinders and 2) the length of their crankshafts. The concurrent existence of several twisting forces of different frequencies (e.g. due to the inertia force harmonics) meant that there were several values of rpm at which the twisting force frequency matched the crankshaft’s natural frequency and caused crankshaft torsional vibration. At worst the result was broken crankshafts, at best it was noisy timing gears, if mounted at the free end of the crankshaft, and roughness at certain rpm.

Much work was carried out in the early days to address the torsional vibration problem and in 1910 Dr Frederick Lanchester patented a vibration damper, for fitting to the free end of the crankshaft. This device was a frictional type of torsional vibration damper, comprising a small flywheel driven by friction material to allow slippage to occur, and hence damping, when torsional crankshaft vibrations arose. Henry Royce (later Sir Henry) evolved a similar device, working independently of Dr Lanchester, and the device was used in Rolls-Royce cars for many years.

Torsional crankshaft vibration was at its most destructive when the rpm in a six-cylinder engine reached a third of the crankshaft’s natural frequency, referred to as third order vibration. Without very effective damping it was dangerous to run engines at this speed. Crankshafts were made stiffer to raise their natural frequency so that third order vibration was removed from the normal operating rpm. Less troublesome vibrations were handled by the vibration damper, or, before dampers were introduced, simply lived with.

When Thornycroft introduced six-cylinder engines during the late ’twenties, the firm did not fit vibration dampers, but placed the timing drive gears at the flywheel end of the engine, rather than the front of the crankshaft where torsional vibration amplitude was greatest. This prevented timing gear rattle at rpm values where crankshaft torsional vibration occurred. However, putting timing gears at the flywheel end of the engine did not prevent crankshaft torsional vibration, and Thornycroft soon fitted vibration dampers to its six-cylinder engines and relocated the timing gears to the front of the crankshaft.

Vacuum servo-assisted braking

Servo-assisted braking systems fitted to Thornycroft lorries amplified the force applied by the driver’s foot on the brake pedal, relieving the driver of having to apply a large pedal force to obtain effective braking. Servo effect was proportional to brake pedal force. In the event of servo failure, the connection between brake pedal and brake drums remained intact, but higher pedal force was needed.

The heart of the vacuum servo system was the servo unit which comprised 1) a vacuum-operated piston to provide lever force at the brake drums, and 2) a compact mechanism which ensured that the piston force was proportional to the brake pedal force. On petrol-powered lorries, the vacuum could be drawn from the engine inlet manifold. However, on diesel lorries an exhauster (vacuum pump) was normally fitted, engine-driven (for example), as inlet manifold depression was insufficient to provide a vacuum. An exhauster could also be used in petrol lorries instead of using inlet manifold vacuum. Sometimes a vacuum reservoir was placed between the vacuum source and servo unit.