It’s now 57 years since the introduction of what would become the most successful Formula 1 engine in history. The Cosworth DFV engine won on its debut at the Dutch Grand Prix in 1967 and would go on to dominate in Formula 1, racking up 155 victories. No other engine has come close. Using notes that were prepared by company founder Keith Duckworth and chief designer Mike Hall, along with feedback from one of Cosworth’s long serving employees Malcolm Tyrrell, we take a look at the birth and technical spec of the early iterations of this incredible power plant.

Cosworth DFV in Lotus 49

Cosworth had already become an established name in the motor racing world with its successful Formula Junior, Formula Three MAE, Racing Lotus Twin Cam and Formula Two SCA engines, based on standard cylinder blocks with bespoke cylinder heads. But then, at the end of 1965, Cosworth embarked on a completely new project that would establish its name in racing for ever. In exchange for £100,000 from the Ford Motor Company, Cosworth would create a four-cylinder Formula Two engine, the FVA (Four Valve Type A), which would then be ‘doubled up’ to create a V8 for Formula One, the DFV (Double Four Valve).

For nine months in 1966, Duckworth entrenched himself in a room at his house, working every day from 9am until past midnight. Once a week he would visit the factory to hand over schemes and drawings, and get feedback on the progress of the manufacture of his designs. Chief designer Mike Hall joined Cosworth from BRM in October 1966, and was instrumental in the detail design of the major components and all auxiliaries, working from Duckworth’s layouts. The engine was completed in April 1967, winning first time out at the Dutch GP in June in Lotus’ new Type 49 chassis.

As per the contract with Ford, Duckworth and his team started with the FVA, incorporating the lessons learnt on this four-cylinder engine into the design of the DFV. Duckworth set a power output target for the FVA of 200 hp, reasoning that the DFV would need 400 hp to be a winning engine. As it turned out, both these performance targets were comfortably exceeded.

Engine Mounting and Layout

The oval shape of the Lotus 49 monocoque meant it would be difficult to extend the sides of the chassis down past the engine, resulting in the dictate from Lotus’ Colin Chapman that the engine should be a stressed member of the chassis from the outset. Part of the suspension would also mount directly to the engine; schemes and handwritten notes still at Cosworth show that the layout of the suspension went back and forth between Duckworth and Chapman before settling on the final positions.

Cosworth DFV in Lotus 49

The engine was bolted to the bottom of the rear of the monocoque bulkhead by a wide-based bracket on the sump, with the intention that these lower mounts would transmit the shear loads between the chassis and engine. Anecdotally, Duckworth would remark that the distance between the two engine mounting bolts at the front of the engine was chosen to match the width of Jim Clark’s posterior! At the top of the engine, the left- and right-hand cam covers were connected to the top corners of the rear bulkhead with thin triangular shaped steel plates, so as to take the tension and compression loads. The plates were also intended to deflect under the thermal expansion of the engine, calculated at the time to be 0.015 in (0.4 mm).

Duckworth was intent on keeping the engine as compact as possible, so to keep the front of engine flat and uncluttered he positioned the water, oil and fuel pumps along the sides of the cylinder block underneath the exhaust pipes. That would also allow for a lower centre of gravity, recorded at 4.6 in (116.4 mm) above the crank centreline. The left- and right-hand pump pulleys were driven at the front of the engine using a Uniroyal toothed rubber belt, driven from a pulley connected to the second compound gear running at half engine speed.

The overall dimensions of the Cosworth DFV engine showed that it was wider and higher than it was long – the height of the engine from the bottom of the sump to the top of the trumpets was 23.3 in (590 mm), the length measured at 21.6 in (550 mm) and the overall width between the extremities of the cam covers was 26.8 in (680 mm). With no clutch or starter motor fitted, the dry weight of the engine came out at 350 lb (159 kg).

Crankcase

With the rules stating that the DFV had to have a capacity of 3 litres, the individual cylinder capacity was reduced from 400 cc on the FVA to 375 cc on the DFV. The DFV copied the FVA’s bore diameter of 3.373 in (85.67 mm), with the smaller cylinder capacity achieved by a shorter stroke of 2.550 in (64.77 mm) to give a total capacity of 2993 cc.

Cosworth DFV Block & Sump

Like most of the other major castings, the cylinder block was created from fully heat-treated 7% silicon LM 25 aluminium. At the time, this material was considered to be the best commercially and readily available alloy in the UK, with the highest proof stress and Brinell hardness, and good casting and machining capabilities.

In the cylinder block were wet liners machined from centrifugally spun castings, made from chrome vanadium alloy iron. They were interfered into the block with two O-rings at the bottom of each liner to seal off the crank case chamber. The top of the flange on the liner had a recess for a Coopers sealing ring.

The sump was made up of box sections that ran from front to rear and across the engine to try to maximise stiffness. The box section at the front also carried water to connect and balance the two water pumps on either side, while the rear box section carried the scavenged oil to an outlet on the left-hand side of the engine.

Water Pumps

The first iteration of DFV engines had a single water pump, mounted on the right hand side, directly in front of the oil scavenge pump. Later DFV iterations featured two water pumps mounted on either side of the engine, so that each bank of the engine was cooled by its own water pump.

These later water pumps used 2.5 in (63.5 mm) diameter centrifugally bladed impellers contained within a volute spiral casing. Each pump had a maximum flow capacity of 45 gallons per minute (204 litres per minute), and the water flow from each pump was sufficient to restrict the temperature rise from the water pump inlet to the engine outlet to 7 C.

Fuel Pressure Pump

A mechanical fuel pressure pump was positioned at the front of the left-hand side auxiliaries. This gear-type pump was capable of delivering 40 gallons per hour (182 litres per hour) at maximum engine revs against a back pressure of 120 psi (8.3 bar). An electrically driven high-pressure pump was activated for starting purposes; this auxiliary pump was then switched off when the engine speed reached 2500 rpm.

The fuel pump was positioned as far out as possible so that it would be cooled by the air stream around the engine, in a bid to keep the fuel cold and prevent fuel vaporisation. A further measure to reduce fuel temperature included isolating the pump from its supporting bracket with a Tufnol insulator.

Oil Pressure Pump

At the rear of the left-hand auxiliaries sat the main oil pump incorporating an integral filter. The oil pump contained a 0.8 in (20.3 mm) wide Hobourn Eaton lobe-type rotor, which at maximum engine revs could displace 11 gallons per minute (50 litres per minute) against a 100 psi (6.9 bar) pressure relief valve setting.

A transfer pipe fed the filtered oil from the pump to the crankcase main oil gallery. From here the oil was fed to the five pairs of main bearings, and then through the cross-drillings in the crankshaft into the eight pairs of big-end bearings. It was estimated that at 11,000 rpm a minimum oil pressure of 68 psi (4.7 bar) was required to force the oil into the centre of the crankshaft due to the centrifugal head generated at this speed.

Oil Scavenge Pump

On the right-hand side of the engine were the scavenge pumps, two pairs of Hobourn Eaton form rotors with a separator between them, drawing oil in from the scavenge chambers in the sump. (Later pumps would feature Roots type rotors.)

Each cylinder head featured two head drains, one at the front and one at the rear, which fed down to the scavenge chambers. Finally, oil from the oil pump pressure relief valve was also fed directly into one of the scavenge chambers.

Early DFV engines suffered serious issues with bad draining of oil from the cylinder heads. The first engines only had a scavenge capacity of twice the oil pressure pump, so the scavenge pumps were unable to handle the oil and the blow-by gases. As a result, the blow-by gases would go up the head drains, preventing oil from coming down the other way, which caused the heads to fill with oil, eventually venting the oil to atmosphere and draining the oil tank.

Unable to increase the size of the head drains, a temporary sliding vane-type pump was fitted to take care of the blow-by. However, this in turned caused more problems, as the conventionally scavenged oil, which measured 11 gallons per minute (50 litres per minute), was now mixed with the equivalent of 40 gallons of air per minute (182 litres per minute). The resulting aerated oil led to the design of a centrifuge which ensured that the mixture returned to the oil tank contained 90% oil and only 10% air. This centrifuge was included in the new design of scavenge pump, which now had a capacity of 55 gallons per minute (250 litres per minute) – five times the capacity of the oil pressure pump.

Geartrain

Behind the front cover was a geartrain consisting of 14 gears. Assembled to the front of the crankshaft was the crank gear, which turned the first compound gear, an assembly of two gears. The smaller of these two gears drove another gear assembly, termed the second compound gear. This assembly was made up of three gears; a large central driven gear sandwiched between two smaller gears. To the left and right hand side of the second compound gear were the first of two head idler gears, driven from the outer gears of the second compound gear. On each bank, the first head idler gear turned the second head idler gear. Finally, the left and right hand second head idler gears drove the respective bank’s inlet and exhaust cam drive gear.

All the gears were made from forged vacuum re-melted EN39B steel blanks, case hardened to a depth of 0.020 in (0.5 mm). Considerable effort in production engineering and quality control was made to ensure that the teeth were accurately ground to ensure concentricity of their pitch circle diameters to the gear bearing bores, resulting in the required backlash and correct involute profiles.

However, despite the great attention to detail given to the design and manufacture of the geartrain, gear problems blighted the first race engines. At debut race of the Cosworth DFV engine, Graham Hill retired with cam gear failure, and broken gear teeth were found in the winning engine of Jim Clark. In addition, the gear failures were compounded by catastrophic valve spring failures when the engines were run at more than 9000 rpm. Cam lobes profiles were redesigned to bring down their maximum torque requirement from 36 lb-ft (49 Nm) to 26 lb-ft (35 Nm); however, even though the valve spring life improved, there were still gear failures.

Using strain gauges, instantaneous stab torques of 300 lb-ft (407 Nm) where recorded – far higher than the original torque calculations used for the gears. What was needed was a way to absorb the shock loading from the camshafts that was destroying the gears.

Cosworth DFV Compliant Compound Gear

In a typical moment of Duckworth ingenuity, the answer came in ‘cushioning’ the second compound gear, which up to this point had been a rigid assembly. The new design of second compound gear contained 12 small quill shafts that allowed the two side gears to rotate over a limited angular displacement relative to one another, thereby storing some of the huge energy from the cam loadings to successfully reduce the loading on the geartrain.

Crankshaft

Although the first crankshafts were machined from billets, Cosworth quickly switched to fully forged blanks supplied by Smith Clayton Forge. The material for the crankshaft forging was heat treated EN40C 3% chrome molybdenum nitriding steel. After the crankshaft was machined, it was nitrided in an ammonia atmosphere at about 500 C, resulting in a hardened case depth of around 0.015 in (0.38 mm).

The main journals had a diameter of 2.375 in (60.325 mm) and the crank pins were the same as the FVA at 1.938 in (49.2 mm) in diameter, with the crankpins arranged to give a flat-plane layout. The crankshaft weighed 32 lb (14.5 kg), coupled to a flywheel weighing 8 lb (3.6 kg).

Calculations made during the design of the crankshaft showed that the maximum load would be on the centre main journal, creating a bearing pressure of 6600 psi (45.5 MPa) at 10,500 rpm. The big-end bearing pressures were estimated to be nearly 8000 psi (55.2 MPa). Both the main journal bearings and the big-end bearings were made by Vandervell, from a steel backing with a bronze intermediate layer and a lead indium overlay.

During the early 1970s a Holset-manufactured crank damper was situated on the nose of the crankshaft to reduce torsional vibration. Analysis by both Holset and Vandervell proved that the principal resonance peaks lay between 8,000 and 11,000 rpm. The largest of these peaks was the eighth harmonic excitation of the first order, which occurred at 8,594 rpm, unfortunately in the middle of the running range of the first engines of between 7,000 and 9,500 rpm. The maximum alternating torque was +/- 2,122 lb-ft (2,877 Nm) on the third crankpin, creating a maximum amplitude of +/- 0.95°. The crank damper was removed on later engines in the mid 1970s when the running range was increased away from this peak to 9,000-10,500 rpm.

The crankshaft proved to be an extremely reliable component, but in 1970 there were a series of widely known crank failures. Investigations pointed to a relatively simple grinding error. The corner radii of the crankpin journals were ground both before and after nitriding, but unfortunately the radius on the pre-grinding wheel was too large, which led to the post-nitride grinding wheel going through the nitrided layer, drastically reducing the life of the crankshaft.

Piston & Rings

DFV piston forging material was chosen to be RR58 aluminium alloy (developed by Rolls-Royce). Although this material had a slightly higher thermal expansion coefficient when compared with the high-silicon alloys used on production engine pistons of that era, it remained consistent from 20-200 C. The skirt profile featured tapering along the length of the skirt combined with ovality around the diameter, to provide a diametral skirt clearance of 0.003 in (0.076 mm).

Cosworth DFV Piston

Duckworth sought to minimise the weight of the cast-iron top compression ring, such that the thickness was only 0.030 in (0.76 mm) thick. This would ensure that the ring would stay seated on the bottom face of the piston groove under deceleration and thereby prevent gas leakage past the ring. For the record, the ring gaps were set at 0.017-0.022 in (0.43-0.56 mm).

The gudgeon pin was made from heat treated EN39, case hardened all over, with an outer diameter of 0.813 in (20.6 mm) and an inner diameter of 0.47 in (11.9 mm).

Connecting Rod

Both the rod and the cap were supplied as separate stamped forgings by Smethwick Drop Forgings, using re-melted EN24 steel. The cap was secured to the rod with a pair of 3/8 in (9.525 mm) 12-point big-end bolts, with location provided by two dowel pins. The small-end bush was a steel-backed bronze bearing supplied by Vandervell, finish-bored and honed after assembly. The rod centre distance was 5.230 in (132.84 mm), as Duckworth tried to keep the rod length as long as possible to reduce the secondary out-of-balance forces inherent in a V8 configuration.

Cylinder Heads

The FVA cylinder head featured an included valve angle of 400. Duckworth reduced this further on the DFV to 32° to give a shallower pent-roof chamber and hence a further reduction in surface area and hence less heat loss.

Both the left- and right-hand heads were machined from a common casting. These heads contained 1.32 in (33.5 mm) inlet valves and 1.14 in (29 mm) exhaust valves, both with 7 mm diameter stems. In the middle of the combustion chamber was a 10 mm spark plug. The valve seats and guides were made from aluminium bronze alloys, with 0.003 in (0.076 mm) interference in the head when cold.

Cosworth DFV Cylinder Head

The shape of the inlet ports was kept as straight as possible by Duckworth, with a diameter of 1.02 in (25.9 mm). The ports were fully machined; straight sections were bored, while curved and flared sections into the throats were copy milled. The exhaust ports were completely curved and so had to be copy milled throughout. After machining, the heads were given to the fabled finishing section, where the ports would be polished using a process known as ‘broddling’ within the organisation. It was not uncommon for the finishers to stamp their initials on the side of their cylinder heads so that they could compare dyno test results with each other.

Duckworth had a very rational approach to port design. “I have never believed that there is any point in having a gas flow rig and measuring the flow,” he once said. “I think it is possible to look at the shape of a hole and decide whether the air would like to go through it or not. A hole that looks nice and smooth and has no projections will generally flow easily. Most people start with something so horrible that to create an improvement should be very simple. I would claim that I could arrive at something close to their results from gas flowing just by putting my finger down the hole and seeing what it feels like.”

Camshafts

The camshafts ran in steel-backed white metal shell bearings, again supplied by Vandervell. The bearings were held between dowelled caps and one-piece cam carriers, which like the heads were made from a common casting. Oil was fed up from the main engine gallery into grooves in the middle bearing pairs and then into the hollow camshafts, where it would be directed through drillings in the other cam journals to lubricate the other bearings. The oil from the cam bearings also splash lubricated the tappets, after initial tests showed that feeding oil through holes in the cam lobes was not necessary.

The selected material for the camshafts was EN16T steel, which was liquid nitrocarburized (Tuftride) all over after machining to provide an anti-friction coating. Tappets were machined from EN40, fully ground all over and lapped on the tappet face. The tappets ran directly in the cam carrier, and were 1.25 in (31.75 mm) diameter by 0.9 in (22.9 mm) long.

Understandably, Duckworth paid a lot of attention to the profile of the cam lobes. The Cosworth DFV engine copied most of Cosworth’s other engines of that era and had a lift of 0.410 in (10.4 mm), and symmetrical valve timing with inlet valve opening at 58° before TDC and closing 82° after TDC, exhaust valve opening 82° before BDC and closing 58° after TDC, giving 116° overlap. During build, the tappet clearances were set to 0.010 in (0.254 mm) on the inlet side and 0.015 in (0.38 mm) on the exhaust side.

The ‘Bomb’

In the centre vee of the Cosworth DFV engine lived what Cosworth termed the ‘bomb’, a set of auxiliaries driven from the second compound gear. Within a magnesium centre casting was a Lucas rotating magnet alternator that produced 10 A at 12 V. Also in this assembly was the Lucas Opus ignition system (Oscillating Pick-up System), a plastic drum rotating at half engine speed into which was moulded eight ferrite rods running against a stationary pick-up.

Within the ignition system was a thyristor speed limiter set to 11,300 rpm. This was a standard Lucas product, which Cosworth would then wire into a rubber-mounted box before subjecting it to numerous rig tests to ensure consistent operation over the required speed range, with an overspeed test to check that the speed limiter operated correctly. The trigger disc for the ignition system was mounted on the nose of the crank.

Finally, also in the centre vee was the fuel injection metering unit, again supplied by Lucas. The unit consisted of a stationary hollow sleeve containing a series of radially drilled holes, some of which would feed fuel in from the high-pressure pump and some which would allow fuel out to the injectors mounted in the inlet trumpets. Within the sleeve ran a rotor that also had a corresponding array of radial holes, at selected angles to give the required timing of fuel delivery. Along the centre of the rotor were fuel metering shuttles that oscillated back and forth. A fuel cam that pivoted at the end of the unit controlled the length of the stroke of these shuttles, thereby determining the amount of fuel being delivered. The cam lever was driven by the throttle slides, so that when the throttles were fully open the shuttles could operate a maximum stroke and hence provide maximum fuel delivery to the injectors.

In some of the early races in the late 1960s the Cosworth DFV engine was plagued with fuel vaporisation issues, especially at races in South Africa such as those held at East London and Kyalami. The solution was the rerouting of the return fuel line from the fuel PRV to help cool the fuel.

The Launch of an Icon

The DFV was unveiled by Walter Hayes, public relations director of Ford, at a function at Ford’s Regent Street showrooms on 25th April 1967. In attendance were Keith Duckworth, Lotus founder Colin Chapman and Lotus driver Graham Hill.

Cosworth DFV Unveiling

Autosport had this to say about the launch of a new Ford Formula 1 engine: “The announcement this week of the Cosworth-designed and developed Formula 1 engine must be a cause of concern for Lotus’ rivals in the Grande Epreuve field. The compact V8 is extremely light and, as it is designed to do the work of the chassis frame at the rear of the car and carry the suspension, the new Lotus that has been designed around the engine can also be expected to be light and compact. With minimum weight, 400 bhp and Jim Clark and Graham Hill as their drivers, Team Lotus and Ford must be very strong contenders for World Championship victory once they have got their new car sorted out.”

Autosport’s words would prove to be very prophetic – Jim Clark won the first race that the new Lotus 49 powered by the DFV entered in 1967, and Graham Hill took the driver world championship the following year.

Lasting Legacy

The DFV would become the most successful engine ever in the history of Formula One. It went on to win 155 Grand Prix races, 12 driver world championships and 10 constructor championships. It spawned several other winning engine types – the DFW (Tasman Series), DFX (CART/IndyCar from 1975), DFL (Group C), DFY (Formula One from 1983), DFZ (Formula One in 1987), DFR (Formula One from 1988) and the DFS (CART/Indycar from 1988). And its DNA could be traced through subsequent Cosworth Formula One engines and even in the current generation of race engines.

Although numerous developments of the DFV were pursued by Cosworth – including shortened stroke, new cam profiles, changes to the ignition system, redesigned pumps and larger valve diameters – its basic layout never changed. Perhaps this is even more remarkable given that the DFV was the first entire engine designed by Cosworth.

The DFV is still very much alive in historic racing. We continue to support customers rebuilding these engines, supplying a wide range of genuine Cosworth parts along with parts that we’ve been able to reverse engineer and develop. You can find these parts in our on-line shop here: https://modatek.co.uk/product-category/dfv-parts/


This feature on the Cosworth DFV is based on an article written by Modatek’s Matt Grant for Race Engine Technology, issue 84. You can purchase a back copy of this publication here: https://www.highpowermedia.com/Product/race-engine-technology-issue-084

Every engine lubrication system needs pressure to move the oil around the myriad of galleries and drillings inside the engine. One or more pumps will provide the pressure, and the most common way to regulate the pressure is with an oil pump PRV (pressure relief valve) like the one shown below.

Oil Pump PRV Assembly

The oil pump is a pulsating heart that continually moves the oil so that it can lubricate running surfaces and remove heat from critical areas. If a typical high performance engine has an oil flow rate of around 60 litres per minute and an oil volume of around 6 litres, that means that in the 30 seconds or so that it’s taken you to read this far, the oil in the lubrication system will have already been pumped around the engine five times.

All engines need at least one oil pressure pump to create enough pressure to push the oil around the oil circuit. In addition, an engine with a dry sump will also have one or more scavenge pumps, whose job is to remove the oil from the various sections of the engine.

Positive-Displacement Pumps

Both oil pressure pumps and scavenge pumps tend to be of the positive-displacement type, whereby the internal elements of the pump move to create a void that opens up and expands, filling the void with oil. The elements continue to move in a manner that then compresses the void, forcing the oil out of the pump and into the oil circuit.

A high performance oil pump normally contains rotating elements such as rotors or gears, and they typically either run inside one another (classed as an internal gear pump) or side by side (referred to as an external gear pump).

These rotating positive-displacement pumps need a rotary drive, which on most race engines comes courtesy of the crankshaft or camshaft, either directly or from gears, chains or belts that in turn are connected to the crankshaft or camshaft. Consequently, the pressure delivered by the pump increases with engine speed.

In the most extreme circumstances, if the pressure in the oil circuit isn’t regulated then it could rise to such a level that could damage the engine. For instance, extreme pressures could rupture the oil filter or blow out any sealing plugs. Most oil circuits will therefore include at least one device that can control and regulate the pressure, to make sure it can’t exceed a predetermined maximum level.

Spring-loaded Piston PRV

The majority of PRVs are constructed from a spring-loaded piston, which moves when the oil pressure in the dead space above the piston reaches a certain level. When the piston moves, it reveals a port that allows the oil to vent out of the pump, thereby capping the oil pressure. While the theory behind such a device is relatively simple, as ever the devil is in the detail.

The operation of the oil pump PRV depends on a correctly defined spring (the one pictured above is our high pressure spring for the Cosworth YB oil pump). Some good old-fashioned engineering equations can give accurate results to define the spring required in the PRV.

Let’s use the following terms:

  • Required PRV opening pressure = Preq
  • Spring force = F
  • Piston radius = r
  • Spring rate = k
  • Length of travel of piston required to open PRV outlet port = x

The pressure on the piston is simply the spring force divided by the piston area, which is:

The spring force can be derived from Hooke’s Law:

Combining these two equations, we get:

However, we must also consider the back-pressure on the other side of the piston, downstream of the PRV. Although small, the back-pressure will affect the movement of the piston and should ideally be less than 10% of the required opening pressure of the PRV.

This back-pressure is a result of a huge range of variables, such as bearing clearances, oil viscosity and temperature, flow passage size and roughness, plus more. It is therefore hard to calculate but can be found from measurements taken when the engine is running.

The back-pressure will have an effect on the required pressure, and hence needs to be included in the calculation. If we term the back-pressure as Pback, then the calculation for the required pressure becomes:

So, if we’ve already decided on the piston radius and the length of travel of the piston (which is normally dictated by the space that is available for the PRV), we can rearrange this equation to choose the correct spring rate that will correspond with the required opening pressure:

Armed with this information, the spring designer can decide on the wire diameter, coil diameter, number of active coils and shear modulus to give the required spring rate, k, from this equation:

where d is the wire diameter, D is the coil mean diameter, N is the number of active coils and G is the shear modulus.

Spring Stress

The spring designer also has to check that the torsional stresses in the spring are within safe levels, so that the spring doesn’t elastically deform or break. FEA is one way to determine the stresses, but again some simple calculations can also be used to good effect.

The torsional stress τ in a spring under load F can be found from the equation:

KW is known as the Wahl factor, and is a corrective factor that takes into account the effect of direct shear and the change in coil curvature, and can be found from this equation:

where C is the spring index:

Checking the stresses in a PRV spring might seem unnecessary, but the effects of a broken PRV spring can be as serious as a failed valve spring. The PRV will continue to operate with a broken spring, but if it opens at a lower pressure than required, the oil supplied to critical components such as the crankshaft bearings and piston squirt jets will be at a lower pressure, and like a broken valve spring, the result can be catastrophic engine failure. Combining the above series of equations into a spreadsheet though can enable a PRV designer to quickly establish the link between required pressure and torsional stress in the spring.

Adjusting the Required Pressure

It is also possible to modify the required pressure of an existing PRV without resorting to changing the spring. If the spring has a fitted length of L1 and the length of the spring is L2 when it is compressed by x, then:

We can see from this equation that we can increase the required pressure by reducing the length of spring when it is compressed (L2). In reality, this can be done during assembly of the PRV by adding one or more shims to one end of the spring.

Oil Pump PRV Reliability

One would think that being constantly flushed with oil, the piston would be free to move up and down the cylindrical bore in the sleeve without any issues. However, one of the biggest problems with spring-loaded piston PRVs is that they can tend to jam. That is especially true if dirt or debris gets trapped in the gap between the piston and the bore.

For this reason, suitable filtration of the oil is vital, and usually the oil pump PRV is located downstream of the oil filter so that it receives the oil in its cleanest state.

Also, the clearance between the piston and the cylinder is kept as low as possible, to keep any contaminants out. It’s not uncommon for the piston and the sleeve to be machined as a matched pair. Normally the piston diameter will be measured, then the bore in the sleeve will be machined to the correct size to give the required clearance for that particular piston.

Some PRVs have clearances as low as 5 microns, so it’s essential that this machining is as accurate as possible. Usually, the piston is ground and the bore in the sleeve is honed, as both of these manufacturing methods result in an extremely low dimensional tolerance of just a few microns and can give exceptional levels of circularity and run-out.

The material for the piston and sleeve tends to be as hard as possible, so that any debris doesn’t scratch the walls of both parts. Also, the leading edge of the piston is normally kept as sharp as possible – even a small chamfer can trap debris that will then find its way into the radial gap between the piston and bore.

Also, it is important that the piston and sleeve are demagnetised, otherwise small magnetic forces can cause the piston to stick.

Oil Pump PRV Bypass Return

The choice of where to route the oil that comes out of the PRV bypass seems to be a matter for debate. The two options are to feed the oil back to the oil pump’s inlet or to ‘dump’ the oil back into the engine, usually into the sump or oil tank. Returning the oil from the PRV to the oil pump inlet is the more popular option on race engines.

One benefit of returning the PRV flow back to the oil pump inlet is that it can help to prevent the onset of cavitation. Briefly, cavitation is the damage caused to a surface by the formation of tiny bubbles in the oil. The bubbles are created when the pressure in the oil drops below the oil’s vapour pressure. The oil will boil, instantaneously creating thousands of these tiny bubbles. When the pressure in the oil rises above the vapour pressure again, the bubbles instantly collapse.

This rapid movement of oil leads to small zones of highly pressurised oil, which when combined with the shockwaves from the collapsing bubbles can result in pitting of any nearby metallic surfaces.

There are a number of tricks that can stop cavitation, and most methods are aimed at increasing the pressure at the inlet to the oil pump. For example, if there is a filter on the pump inlet then the mesh size could be increased, or the inlet port to the oil pump could be contoured to help the oil flow more easily into the pump.

Oil Pressure Requirement

I have designed numerous oil pump PRVs in my time, and I’ve found that the most difficult part of the process is actually deciding the pressure the PRV should open up at. Ask a group of engine designers how much oil pressure an engine needs and you’ll probably receive a number of conflicting answers.

Some will say the pressure needs to be high enough to keep highly loaded bearings lubricated or to feed the piston squirt jets. However, others will say the demands of creating too much pressure can increase the parasitic power losses, so it needs to be as low as possible.

In truth, the exact oil pressure required will normally be decided only after multiple tests, either in the car or on a dynamometer. It is important therefore to make sure the designed PRV has room for adjustment, as mentioned already by the use of shims or via an external adjustment device.

An old rule of thumb used to be that an engine needs 10 psi of pressure for every 1000 rpm of engine speed, so for example an engine that revs to 8000 rpm will need 80 psi. In reality, that is probably an over-cautious estimate for modern race engines. Developments in both lubrication and bearing technology mean that the higher grades of oils used in a race engine can withstand more extreme pressures in the plain bearings of the crankshaft (the area in which the oil is typically the most stressed), and enhanced additives allow the oil to behave better for longer.

One train of thought is that the required oil pressure is closely related to the clearance of the crankshaft bearings combined with the viscosity of the oil, because oil pressure can drop if the viscosity is reduced or if the bearing clearance is opened up. Given that reduced bearing clearances can have a positive effect on power, some engine builders will offset the increase in required pressure that might arise from reducing bearing clearance by running lower viscosity oils.

In truth, one has to consider the various sources of pressure drop along the entire lubrication circuit when specifying the oil pump’s required pressure. Some pump designers will group these sources together to come up with an ‘effective orifice area’, which is the equivalent flow area of all of the holes and gaps that the oil has to flow through. It will be a combination of the bearing clearances, plus clearances to other mating parts such as camshaft followers along with small holes such as those in squirt jets.

Broadly speaking, oil pressure is proportional to both the effective orifice area and the oil viscosity: the oil pressure rises if either the effective orifice area or the oil viscosity is increased. This can be observed when starting a cold engine – the lower temperature means the clearances are small and the oil is thicker. Both of these effects combine to give higher oil pressure.

As the engine gets hotter, the clearances in the engine begin to open up, and the oil gets thinner. The increase in both the effective orifice area and the oil viscosity will result in a drop in oil pressure, so it is vital to make sure that the chosen oil pressure requirement is optimised for the required range of temperatures the engine will experience.


This feature on oil pump PRVs is based on an article written by Modatek’s Matt Grant for Race Engine Technology, issue 136. You can purchase a back copy of this publication here: https://www.highpowermedia.com/Product/race-engine-technology-issue-136

OK, so this might not be the most exciting of subjects, but understanding the conventions of Cosworth part numbers can save a lot of time and effort. If you’re trying to identify a part then the part number that might be marked on it can give you several clues as to what engine it came from.

To understand the history of Cosworth part numbers, let’s go back 65 years to when Cosworth was first formed. Cosworth’s first foray into engines centred on modifications to the Ford Anglia engine, producing components like camshafts. They soon progressed to building cylinder head assemblies, followed by complete engines that were designated Mk I, Mk II, Mk III, Mk IV, etc, followed by engines such as the MAE, SCA, FVA and DFV.

One of Cosworth’s early strengths was that every part was identical from batch to batch, so it was easy to swap parts between engines. There were minimal modifications required prior to fitting each part, and this made engine build a lot quicker and far more reliable. They achieved this through very detailed drawings and technical specifications that would give the manufacturer all the information they needed to produce the part. Cosworth would then book the part into stores when it was either manufactured in-house or bought in. Then all Cosworth had to do when building the engine was to book out the required parts and assemble them into the engine.

DFV Parts List

The production and tracking of components for these engines necessitated a system to keep track of the parts that would be required – the parts list. Hence each part needed its own unique part number, one that could be used to track the life of the part through design, manufacture and assembly. Cosworth quickly realised that random part numbers wouldn’t work, and that they would need to have a system to follow to generate each part number.

The Project Codes

The system that Cosworth devised was fairly simple (remember KISS – Keep it Simple, Stupid!). Every part number consisted of two letters that designated the project code followed by four numbers, for example YB1429. With the amount of engine projects rapidly expanding as the company developed, this system meant that one could instantly identify which engine the part was intended for.

There were over 50 different project codes, and some of these never saw the light of day, but here are some of the more common ones, arranged in alphabetical order:

Project CodeEngineYear
BABD series1969
CACA series (Formula 1)2006
CKCK series (Ford CR series Formula 1)1999
DADFV1967
DLDFL1981
DXDFX1986
DYDFY1982
DZDFZ1987
FAFVA series1966
HBHB series (Formula 1)1989
JDJD series (Ford Zetec-R Formula 1)1996
SASCA series1964
TALotus Twincam1963
TJTJ series (Ford CR & RS series Formula 1)2003
VJVJ series (Ford Zetec-R Formula 1)1998
WAWA series (Mercedes-Benz)1984
XBXB series (Indy Car)1992
XDXD series (Indy Car)1996
YBYB series1984
YDDuratec2004

Cosworth also used other project codes in their part numbers, especially for parts that might be used on one or more projects. These included:

PP/PR – ‘proprietary parts’, these were for parts that were bought in

LL – ‘liner length’, these were bought in parts that were supplied by length, such as O-ring cords

DE – electronics parts

PA – this code was used for pistons

KK – these were kits of parts, like piston rings

The codes were supposed to be unique for each project, but occasionally the same code was used for two different projects. For example, CA was meant to be the code for the Cosworth 4WD Formula 1 project in 1969, but it was also used for the V8 Formula 1 engine in 2006.

The Casting Codes

Cosworth also had another set of project codes, called the ‘casting’ code, that were specifically for parts like castings, forgings and billets. Again, these codes were linked to the engine project, and Cosworth instigated a clever way of defining the casting code. The first letter was the same as that for the project code, and for the second letter, just go 13 letters along in the alphabet. So, for example, the casting code for YB was YN.

There were some exceptions, and there was also a rule that letters like I and O had to be skipped as they could be confused with numbers, but on the whole the casting codes followed this pattern.

This brings us to an important point. Many people incorrectly identify a part by the casting number, but this only identifies the part when it is in its part-finished form. For example, in the photo below, YN0627 is the part number of the YB head casting, not the finished machined component.

YB Cylinder Head

There will be a number of different types of cylinder heads that are machined from the YN0627 casting, so knowing just the casting number doesn’t completely identify the head. Here are the part numbers for the heads for the different types of YB engine:

Engine TypeCylinder Head Part Number
YBBYB0935
YBCYB0567
YBDYB0937
YBFYB0528
YBGYB0643
YBJYB0643
YBSYB0643
YBTYB0643
YBPYB1043
YBMYB0977

Another example of a common misconception comes with pistons, which had a forging code of PM. Cosworth would imprint the forging part number into the forge tool, and this would be visible on the piston. But the PM part number only referred to the part number of the forging, not that of the machined piston. Given that the same forging could be used for a variety of different pistons, it is the finished part number that is required.

The Four Numbers

As mentioned, the part number consisted of two letters followed by four numbers. For most projects, four numbers would be enough, as it would be extremely unusual for a project to need more than 9,999 part numbers.

At first these four numbers were sequential, starting with 0001. Usually each project had its own folder that listed these numbers so that there could be no duplication. Over the years, Cosworth started to instill some ‘intelligence’ into these four numbers. The four numbers would begin with an 8 for assemblies, and numbers beginning with 05 were reserved for schematic drawings.

At one point Cosworth also introduced a rule that said that the last number would odd for left hand components and even for right hand components.

Certain projects stipulated more rules for the four numbers, such as reserving 0001 for the cylinder block, 0002 for the LH cylinder head, 0003 for the RH cylinder head and 0010 for the crankshaft. However, this rule was fairly short-lived.

YB1429 Head Gasket Drawing

Cosworth used the part number on all documentation, including drawings (as per the example of a drawing excerpt above for our YB1429 WRC head gasket), purchase orders and invoices.

SAP & Sequential Part Numbers

These fairly simple rules for defining the part number with the project code and four numbers ran fairly smoothly for a number of decades. However, all this came to an end when Cosworth launched SAP at the beginning of 2007 as its new ERP (enterprise resource planning) system.

SAP was essentially a giant database that contained virtually all of the company’s records. It enabled Cosworth to be able to keep track of the entire life of a component, from design through to manufacture, assembly and usage. But one of the problems of SAP was that it was no longer possible to allocate part numbers that followed Cosworth’s rules. (Actually, this proved not to be true, but by the time a solution had been found, it was too late.)

Instead, the company switched to sequential numbers starting at 20000000. There was a central computerised database that would supply the next number available, but without the intelligence that the old part numbering had provided.

To make matters more confusing, for the first few months Cosworth used numbers starting at 10000000 to identify raw materials and kits. You’ll notice that some of our piston sets follow this pattern, such as 10001487 for our BDG Hoyle piston sets.

BDG Hoyle Piston Label

If you’ve survived to the end of this article, then congratulations. It is definitely not the most riveting of subjects, but will hopefully help to shed some light on how to identify Cosworth components.

The humble valve spring might appear to be a relatively simple piece of engineering, but in reality it can be the most highly stressed component in a race engine. The move to faster engine speeds and more aggressive cam profiles can often result in terminal failure of the springs leading to catastrophic engine failure, prompting spring replacement earlier than preferred. Engine builders and manufacturers are therefore focusing more and more efforts on ways to increase the life of the spring to avoid costly engine rebuilds.

Cosworth DR4601 Valve Springs

In essence, the prime function of the spring is to provide a force that will keep the reciprocating movement of the poppet valve under control throughout the entire cycle of the engine and at all operating speeds. Loss of this valve control can lead to valve-to-piston contact, extreme loading of the seat in the cylinder head, bouncing of the valve on the seat and damage to the tip of the valve.

A compression spring like those found in engines provides a reactive force when its length is reduced, predominantly owing to a twisting motion of the coiled wire. Under just a static load, it can be assumed that the load in each coil is identical, and if the coil geometry is uniform along the spring axis then the stiffness of each coil is also identical. The highest stress will occur on the inside diameter of the spring, which is where one could expect failure to originate from.

Spring Surge

Of course, when the engine is running, the linear motion of the valve imparted by the rotation of the cam lobe causes a continuous compression and expansion of the spring. As a result, the dynamic loading on the spring has to include the inertia of the spring, which is not considered under static loading. It is the dynamic loading from the inertia that will result in the most common of spring issues, namely spring surge.

Spring surge can be described as vibration of the spring that occurs at a harmonic of the spring’s natural frequency. When describing surge, the movement of each particle of the spring needs to be considered, and in a way this movement of each coil can be seen visually with a child’s Slinky toy spring.

During the initial opening phase of the lift curve, the spring is compressed and the spring coils accelerate. The uppermost spring at the camshaft end will see the entire inertia of the spring, but each successive coil sees less inertia loading thanks to the lower mass below it and lower acceleration due to a smaller deflection. As a result, as we move down the spring away from the camshaft, each coil accelerates and moves at lower values than the one above it. This begins the first compressive wave of the coils, with the camshaft end closing up quickest.

At around the mid-point of the travel of the valve, its acceleration is zero, and at this instant the coils all move at the same speed. Then, as the valve begins to decelerate, the coil furthest from the cam starts to close up more than those above it, creating a compression wave going in the reverse direction. It is this continual cycle of the compression wave that creates a vibration in the spring and is referred to as surge.

Preventing Surge

Whilst there are numerous tricks to reduce or even eliminate surge, the most obvious one is to create a spring with a natural frequency well outside of the running range. The movement of the valve is dictated by the profile of the cam lobe, and can be mathematically broken down into a series of sinusoidal curves with Fourier analysis, from which the harmonics of the profile can be derived, which are expressed as multiples of the camshaft rotational speed.

When one of the harmonics coincides with the spring’s natural frequency, the effects of surge will be pronounced and can result in the compression wave spiralling out of control, leading to loss of contact between the spring and the retainer and spring seat at either end, plus the build-up of excessive stresses in the coils. As the amplitude of the lower harmonics are larger than those of the higher harmonics, some spring designers will recommend that the natural frequency of the spring is at least eight times the frequency of spring operation, whilst some technical publications quote 15-20 times.

Wire Geometry

The cross-section of the wire is usually circular or ovate; the latter term actually means egg-shaped, but in the case of springs this can also be any elliptical shape made up from a number of radii, and can be either symmetrical or non-symmetrical. An ovate spring will typically have the major axis perpendicular to the spring axis, which helps to reduce the stress on the inner diameter as the maximum area of the wire is at the point of highest stress, and can lead to a shorter spring length owing to the wire being slightly flattened.

However, some of the higher grades of steel do not work well with the dies used to make the ovate wire shape, as the additional carbon can extrude the special dies needed to form the more complex cross-section. Consequently, the round wire can be made with more carbon content than an ovate wire. It is also harder to control the orientation of the ovate profile when coiling the spring, as it will have a tendency to twist down the length of the spring during coiling. In fact, changing the external shape of the spring can have a far stronger effect on the life of the spring than using an ovate cross-section.

Spring Shape

Nowadays there are a myriad of options available for the external shape of the spring’s helical coils, although they can be broken down loosely into three categories: straight, conical and beehive.

Conical and beehive springs are termed as progressive, as the stiffness will vary with length. This is also a case for springs with an unequal pitch between the coils along the length of the spring. In progressive springs, each coil has a different stiffness, which means that when the spring is compressed, the coils with a lower stiffness will deform more than those with a higher stiffness. Eventually the less stiff coils become coil-bound (where adjacent coils come into contact), which reduces the number of active coils available as the load is increased, increasing the overall stiffness of the spring in a progressive manner. As the stiffness is varying, so too is the natural frequency, hence in theory progressive springs are less susceptible or even immune to the problem of spring surge.

Both beehive and conical springs also have an advantage over a straight spring in that the retainer can be smaller, which in turn can lead to reduced valvetrain reciprocating mass. The springs themselves can also be lighter, and so the inertia of the upper coils will be lower. One final observation made by a spring supplier is that conical springs can also have a natural alignment action that is very beneficial for very small valve stem diameters.

Nested Springs

Another solution to surge is to use nested springs, where two or three springs are used in parallel, with smaller springs packaged inside larger ones. As with progressive springs, the stiffness of the combined spring varies with length, and each spring will have a different natural frequency, again helping to avoid surge.

Cosworth DR4601 Valve Spring

The outer diameter of the inner spring and the inner diameter of the outer spring are usually chosen so as to create a small amount of interference between the two springs (as is the case with our DFV valve springs). It is essential that the direction of windings is different between the springs, otherwise they will get caught up in one another. The interference will provide a means of damping, allowing unwanted energy to be converted to heat from the friction between the two springs.

When run for extended periods, this interference will of course wear the surfaces of the springs and reduce the life of the nested spring assembly. Owing to titanium’s inherent nature to gall (the macroscopic transfer of material between metallic surfaces) when in contact with other titanium surfaces, titanium nested springs cannot be run with interference.

Where the designer wants to avoid contact between the inner and outer springs, stepped retainers and spring seats can be used. This will separate the two springs to eliminate friction and hence damage to the surfaces.

The size of the spring is determined by multiple factors: the cam profile and associated lift, acceleration rates and opening, closing, flank, nose and seating velocities; valve train masses; the moment of inertia of the rocker (if present); the operating speed of the engine. All of these parameters will give an idea of the required amount of spring travel, spring force and spring rate that is going to be required to control the valve at speed.

Materials

Because of the high stresses that the spring will experience, special care has to be paid to the cleanliness of the raw material, the surface finish and any methods that can be used to leave compressive residual stresses in the surface that would counteract the stresses from running.

When choosing a material, spring manufacturers will look at its torsional modulus of rigidity and torsional yield strength, as well as the more practical requirements such as cost and availability. Most compression springs are made from steel or titanium drawn wire, with the former being more common.

Cosworth PR8121 BD Valve Springs

The actual chemical composition of the steel alloys used by spring manufacturers remains a closely guarded secret, which is understandable given the intense amount of research and testing they carry out to find the perfect mixture of elements. However, what is known is that most steel alloys used in spring manufacture will contain chromium and silicon. Vanadium is also included at small levels to increase the strength of the material, along with manganese, molybdenum and nickel in some cases.

Titanium springs are usually made from Beta-C and LCB (Low Cost Beta) titanium alloys. Titanium can offer the benefit of lower density and higher strength, plus better resistance to corrosion when compared with steel. However, its cost can often mean that that titanium springs are about five times more expensive than their steel counterparts. Also, some of the steel alloys used nowadays have proven to have higher fatigue limits than the titanium alloys available.

Manufacturing

The manufacture of a steel valve spring starts with the material being rolled into rods by a specialist steel mill, that is capable of producing the level of cleanliness required for racing.

Some steel mills will take extra steps to ensure that any inclusions in the microstructure are pushed into the centre of the wire, where the stresses during running will be lower. Also, great care is taken to make the inclusions smaller (it is the inclusions that can make the wire brittle), giving rise to the term ‘superclean’ chrome-silicon as used by some steel mills.

The wire is drawn down to size by pulling the rod through a series of dies, all the while being tested and scrutinised in line with the manufacturer’s quality standards. Attention is paid to tensile strength, surface inspection and chemical analysis, and eddy-current testing is used to verify the surface (an electromagnetic field is created around the wire to allow for microscopic identification of any surface defects such as pitting, cracks and corrosion).

Some spring manufacturers will also use scanning electron microscopes (SEMs) and X-ray diffraction machines (XRDs) to check for material integrity. An SEM allows metallurgists to view the surface topography and composition, while an XRD can measure compressive residual stresses. Such methods can be incorporated into various stages of the manufacturing route of the spring to ensure no degradation in quality.

The wire is then shaped into the designed helical coil pattern using CNC machines to control the winding of the wire onto mandrels, with the wire being either hot or cold. The use of such complex machines allows for better repeatability between batches of springs and improved accuracy in the pitch and diameter of the coils. Next, the spring has to be stress relieved. The coiling stage permanently deforms the wire, creating harmful residual stresses, and so a heat treatment operation at a relatively low temperature is needed to remove them. Note that coiling is more complicated when the wire’s cross-section is not circular.

With the shape of the spring now complete, attention is then paid to the ends of the spring. It is imperative that the end faces of the spring are perpendicular to the axis of the spring and parallel to each other so that the forces will be evenly distributed in the valve stem. As a result, the ends are ground. As the grinding can leave sharp edges, a finishing step is carried out to remove uneven areas on the surfaces of the ends. Without this final operation, the edges could break away into the cylinder head chamber or dig into the retainer and spring seat, creating fatigue crack initiation sites.

Surface Treatments

At this stage, there are numerous processes that spring manufacturers will carry out to increase the life of the spring. We will look here at shot peening, nitriding, polishing and cryogenic treatment, but there are numerous other techniques that manufacturers are less willing to reveal, for obvious reasons.

Owing to its cost-effectiveness and practicality, shot peening is a relatively common technique to impart a compressive stress in the surface. Here, small spherical beads made from steel, glass or ceramic are fired onto the faces of the spring in a controlled manner. The impact of each bead will create a dimple in the surface, stretching it, and below the dimple the movement of the material creates the compressive stresses required.

There are three parameters that can be varied to alter the magnitude and depth of the compressive stress – bead size, intensity and coverage. In general smaller size beads will yield a more polished surface. The intensity is the amount of energy used to project the beads, while the coverage is the amount of area that is impacted by the beads (note that this is always more than 100%). All these variables will depend on the material of the spring and any subsequent processes.

Nitriding is a heat treatment procedure that will diffuse nitrogen into the surface of the spring to give a case-hardened surface and can also impart a compressive stress into the surface. A harder surface is especially useful in a nested spring design, where there is interference between the springs.

In gas nitriding, the spring is placed in an oven at temperatures of about 500°C for a period of time while ammonia is flowed around the spring’s surface. The alternative to gas nitriding is plasma nitriding. Whereas gas nitriding relies on a high temperature to create a reaction with the surrounding gases, plasma nitriding uses intense electric fields to create ionised molecules of the gas (usually nitrogen) around the spring’s surface. Note that if a spring is to be both shot-peened and nitrided, the nitriding step is carried out first, otherwise the high temperatures during nitriding would relieve the compressive stresses induced at the shot-peening stage. The added advantage of nitriding first is that the substrate is harder, so the compressive stress from shot peening is increased.

One or more refinement procedures are also usually carried out to remove any remaining surface defects and imperfections, both between certain operations and at the end of manufacture. Electro-polishing is one such method that has been proven to be beneficial, although it can lead to hydrogen embrittlement, to which the high-strength alloys can be susceptible. However, a combination of chemical and mechanical isotrope finishing is becoming more common, and this creates a polished mirror-like surface without the issues seen with electro-polishing. Some manufacturers will go even further after polishing by adding a final peening operation with minute beads (often referred to as micro-peening or nano-peening).

A final operation carried out after all or some of the above is to pre-set the spring. Here, a relatively large load is applied to the spring, such that while the centre of the wire is elastically deformed, the surface of the wire undergoes plastic deformation. This procedure will set the free length of the spring, as the plastic deformation means that the spring will not return to its original length. Pre-set springs are less likely to relax over time, and if the pre-setting is carried out at a controlled elevated temperature than the spring will be more capable of withstanding service in hot environments too.

Installation

Despite every effort in design and manufacture to increase the reliability of valve springs, spring suppliers see a surprising number of failures due to improper installation. One of the more common issues they see is incorrect design of the retainer. The spring needs to be correctly contained in the retainer to stop it from excessive lateral movement, but not overly constrained such that it is forced into the retainer.

Cosworth DFV Valve Spring, Retainer & Seat

Handling the springs also has to be done with care to avoid damaging the surface. They should never be placed in a vice or pliers, and plastic tooling should be used when separating interference-fit nested springs. Also, springs that have been delivered with a rust preventative coating should not be cleaned with acidic or evaporative cleaners, as this can cause rapid drying and promote the formation of rust on the surface of steel springs.

A static spring testing machine can be used when selecting and fitting valve springs to confirm the rate of the load versus deflection; such machines can detect the onset of binding of the coils.

Summary

The life and maximum operating envelope of many race engines is restricted by the valve spring. While it is possible to extend the life of the spring by reducing engine speed or compromising on cam profiles to lower the acceleration of the valve, there are numerous methods available in the design and manufacture of the spring that should be considered – wire cross-section, the geometry of the helical shape and a combination of nested springs can all be exploited to reduce or even eliminate certain failure modes. Numerous manufacturing processes exist that will create beneficial compressive stresses at the surface.


This feature on valve springs is based on an article written by Modatek’s Matt Grant for Race Engine Technology, issue 89. You can purchase a back copy of this publication here: https://www.highpowermedia.com/Product/race-engine-technology-issue-089

Cosworth CA and DFV Formula 1 Engines
Cosworth CA and DFV Formula 1 Engines

The Cosworth DFV and the CA engines represent bookends of nearly half-a-century of normally aspirated Formula One engines. These two engines allow for a discussion of how Cosworth was able to double both engine speed and power output, as well as guarantee a staggering threefold increase in engine life. In this special tech feature we’ll compare these two engines, taking a closer look at some of the critical components.

An Icon in F1 History – the DFV

Very few companies can survive more than 60 years of trading and still follow the founders’ original mission intent. In the cut-throat world of motor racing this is an exceptionally rare feat, and one only has to look at a race report from half-a-century ago to see that most of the teams and engine suppliers of the time have since disappeared, falling foul of financial setbacks, changes in regulations and necessary diversifications away from motorsport.

Yet one company has flourished since its inception in 1958 and can still claim to be true to its roots. That company is Cosworth, built on the foundation that “it must be possible to make an interesting living messing around with racing cars and engines”, as decreed by founders Keith Duckworth and Mike Costin.

Cosworth DFV Engine Cutaway

In 1967 Cosworth launched the DFV into Formula 1, initially for Team Lotus, and it instantly became a race winner. The following year the DFV was made available to other teams, and it went on to become the most successful engine ever in the history of Formula 1, scooping 155 wins, 12 drivers titles and 10 constructors titles.

The first DFVs might only have been able to delivery just over 400 bhp and reach 9,000 rpm to start with, but as each failure mode was methodically overcome (starting with the valve spring, then torsional gear drive problems), power and speeds gradually rose. By the time the DFV had finished active service in 1985, peak speed had topped 11,000 rpm and power was in excess of 500 bhp.

Working in conjunction with Lotus, the DFV was the first successful Formula One engine designed to be a fully structural member of the chassis, with mounts on the heads, cylinder block and sump connected to the chassis and gearbox bulkheads.

The DFV had a V8 configuration, with the two banks separated by a vee angle of 90 degrees, and for various reasons had the air intake to the two cylinder heads in the centre vee and the exhausts on the outside of each bank, running down either side of the engine below the heads.

The auxiliary water, oil and scavenge pumps were housed on either side of the cylinder block, tucked away underneath the exhausts and driven by a belt that in turn was driven by the nose of the crankshaft. The drive to the camshaft gears was via a series of gears at the front of the engine, again driven from the crankshaft nose. In the centre vee sat the alternator and fuel pump assemblies.

The Need for Speed – the CA

Some 40 years later, Cosworth launched what would turn out to be their final V8 Formula 1 engine, codenamed the CA. Formula 1 engine regulations had gone through many changes since the advent of the DFV, with brief interludes of turbocharging in the mid-80s, followed by a return to normally aspirated 3.5 litre engines up until the end of 1994, and then 3 litre engines in various configurations.

Cosworth CA Engine Cutaway

When the CA was first launched in 2006, and with no regulations capping speed, a mind-boggling and class-leading 20,000 rpm was attainable in qualifying, and by the end of that same season the CA could run up to that speed over an entire race distance. Frustratingly though, regulations aimed at reducing soaring development costs had capped the maximum engine speed to 18,000 rpm when Cosworth returned to Formula One with a modified CA in 2010.

Even so, peak power of 780 bhp was attainable, almost twice that of the DFV when it was first launched, despite a reduction of 14% capacity. Perhaps more impressively, there was a huge jump in engine mileage. The DFV competed at a time of unlimited engine changes, and hence only had to be capable of completing one race distance. Fast forward to the 2010-2013 era, and drivers were only allowed 8 engines for the entire season, which meant that the CA had to be capable of completing around 1,500 miles between rebuilds.

Comparing the DFV and CA

What is perhaps surprising is that the CA’s overall architecture was almost identical to the DFV’s, save for the removal of belt drives and the repositioning of the alternator and fuel pump (in subsequent Cosworth Formula One engines the alternator was relocated to the back of the left-hand auxiliaries, while the fuel pump ended up submerged in the car’s fuel tank and driven by a quill shaft from one of the drive gears on the front of the engine). While Duckworth would certainly never claim to have pioneered this layout, it is interesting to note that most Formula One engine manufacturers later followed a similar approach.

A comparison of DFV and CA engine weights wouldn’t necessarily be fair, as in addition to the reduction in capacity, the CA’s weight was mandated by the regulations, which specified a minimum dry weight of 95 kg. However, it is worth noting that the corresponding weight of the DFV would have been around 168 kg when it was first launched.

The regulations that governed the design of the CA also defined a minimum height of the centre of gravity from the bottom of the sump at 165 mm. As this figure was easily achievable on the CA, any weight-saving requirements were rendered unnecessary, hence extra material in the heads and cam covers could be used to improve the engine’s overall stiffness.

Cylinder Block Comparison

Cosworth CA and DFV Cylinder Blocks
Cosworth CA and DFV Cylinder Blocks

The most striking difference between the cylinder blocks is in their respective sizes. The DFV block stands almost twice as high as that of the CA, due in part to the height of the sump. The distance from the crank centreline to the bottom of the sump for the DFV was more than 133 mm; for the CA that fell to just 58 mm (the minimum allowed by the regulations).

Some of this reduction was made possible by gradually miniaturising the bottom-end geometry over successive engine designs, such as shrinking the crank counterweights following the shift to bolt-on tungsten weights. Of equal significance was lowering the piston stroke (64.77 mm for the DFV versus 39.77 mm for the CA) thanks to an increase in cylinder bore size and a reduction in capacity.

There was a marked increase in bore size from the DFV to the CA, 85.67 mm for the DFV compared with the CA’s 98.0 mm, as dictated by the regulations. Despite this, the CA block is slightly shorter in length, as both engines have almost the same bore spacing. The distance between the walls of the bores could be reduced on the CA because it ran with coated parent metal bores, whereas the DFV block was an open-deck variant containing cast-iron cylinder liners.

Both blocks were cast from aluminium alloys, LM25 TF for the DFV and a similar in-house derived aluminium alloy for the CA. Cosworth actually experimented with magnesium cast blocks and cylinder heads for the DFV during the 1970s, but that was soon abandoned owing to the increased complications that came when using magnesium. In essence, the magnesium blocks and heads were problematic because of the large difference in thermal expansion coefficient values for the magnesium material, the steel main bearings and the nickel-aluminium bronze alloy valve seats.

Cylinder Head Comparison

Cosworth CA and DFV Cylinder Heads
Cosworth CA and DFV Cylinder Heads

Duckworth famously remarked that the DFV was the first race engine to incorporate a narrow included valve angle (the angle between the inlet and exhaust valves). At the time, rival engines had valve angles of around 60 degrees, but Duckworth sought to reduce this to give a shallower pent-roof chamber. The DFV was designed with an included valve angle of 32 degrees – compare this then to only 18 degrees on the CA.

Actually, the CA had compound valve angles, with both the inlet and exhaust valves also inclined 6 degrees apart along the crankshaft axis. The switch to a compound valve angle improved the shape of the combustion chamber on the CA, plus an opportunity to make a small increase in inlet and exhaust valve diameters.

Like the DFV, the CA cylinder head featured a separate cam carrier – in the case of the CA, this was necessary to be able to package the compound valve angles. The CA heads featured a pneumatic valve return system instead of conventional wire springs, which allowed the engine to run at such high speeds.

Piston Comparison

Cosworth CA and DFV Pistons
Cosworth CA and DFV Pistons

Thanks to the tightening up of the included valve angles in the cylinder head, the CA piston crown didn’t need deep valve pockets, as seen on the DFV. Instead, the crown was almost flat, with shallow pockets to clear the valve heads. The lack of sharp edges and pockets in the crown had a huge beneficial effect on good flame propagation and the elimination of detonation.

Piston material was largely unchanged from the DFV to the CA, thanks to the CA’s regulations banning the use of exotic materials such as aluminium beryllium and metal matrix composites. The DFV piston forging used a grade of RR58 aluminium alloy that had originally been developed by Rolls-Royce. By the time the CA was designed, Cosworth had already defined a confidential proprietary blend of aluminium alloy.

Another change that was quite noticeable when comparing the two pistons was the undercrown design, which on the CA was an elaborate arrangement of highly polished ribs and buttresses, optimised using various design techniques. The net effect was a piston that, although bigger on bore size, was almost half the weight of that from the first of the DFVs.

Cosworth Formula 1 CA Piston

One other note was the amount of oil cooling supplied to the CA piston, via an array of squirt jets fed from the main oil gallery in the block. The correct cooling of the piston was a critical factor in achieving the required durability while also withstanding increased gas pressures and speeds.

DFV & CA Continuation

Whilst both engines might have ceased active service, they can still be heard roaring around race tracks today. The DFV is a popular engine for historic Formula 1 race categories, and some of the Formula 1 cars from the end of the V8 era are still being run by privateers with power from the CA.

Modatek actively supports customers with rebuilds of both engines, supplying genuine parts such as pistons, bearings, seals and countless other critical components. We even have an on-line shop for DFV parts.

We also sell race-used Cosworth CA engine parts in our Memorabilia section of our on-line shop.


This feature on the Cosworth DFV and CA engines is based on an article written by Modatek’s Matt Grant for Race Engine Technology, issue 100. You can purchase a back copy of this publication here: https://www.highpowermedia.com/Product/race-engine-technology-issue-100

There is nothing ‘plain’ about a plain bearing, especially when it comes to fluid film journal bearings, like those found on the crankpin and mains journals of the crankshaft. There have been continuous advances in crank bearing material, which enable bearings to last longer between rebuilds.

The first examples of journal bearings were made from Babbitt metal (sometimes also referred to as white metal), named after Isaac Babbitt who in 1839 developed an alloy to be used as a bearing surface in plain bearings. The exact composition was kept a closely guarded secret for a number of years, but was eventually disclosed as an alloy of lead, tin, copper and antimony. These days we tend to refer to any lead- or tin-based alloy material as a Babbitt metal.

Babbitt metals tend to be extremely soft, especially when compared to the hardened crankshaft journals. However, the composition of a Babbitt metal is a matrix of small, hard crystals contained in a softer metal. The aim is that, as the bearing wears, the softer metal yields and creates routes for the oil to pass through, improving lubrication and giving the bearing some degree of conformability.

Originally, a thin layer of Babbitt metal would be applied directly to the bore of the substrate material, but the need to periodically refurbish this surface led to the introduction of interchangeable steel or bronze shells onto which the Babbitt metal could be applied.

Thin Walled Bearings

The most famous example of the replaceable bearing shell that was first developed for high-performance engines was the Vandervell ‘thin-walled’ shell bearing. As the name suggests, the shell comprised a thin strip of steel that was rolled to create a semicircular shell, and the bearing material could then be coated onto the inner diameter, creating what was termed a bimetallic bearing. The steel backing was made from a high-strength steel alloy that allowed the bearing to be interfered into the rod and cap halves without yielding.

Today’s bearings still rely on a high-strength steel alloy backing in most cases, but here the similarities with their distant ancestors ends. Bearing manufacturers discovered that a steel-backed bearing combined with a thin Babbitt metal layer was prone to wear, reducing the time between bearing change intervals. If they wanted their products to last longer and be more reliable, they had to find a way of adding strength and wear resistance into the bearing. That thinking led to the introduction of a multi-layered bearing, which could provide the compromises needed between all the functions the bearing had to provide.

Bimetallic bearings were therefore usurped by trimetallic bearings, which are composed of a backing, a substrate (or lining) and an overlay. The substrate layer gave the bearing its load-carrying capability and provided resistance to wear and cavitation.

Bearing Material

The most common material for the substrate layer is a copper and lead alloy, which replaced the previous versions of tin-based Babbitt metals a number of years ago. Such alloys tend to consist of 20-40% lead, with the rest made up of copper and sometimes small amounts of tin, silver or nickel.

Cosworth GB0047 Rod Bearing

Vandervell (now part of the MAHLE group) created its own unique specifications for crank bearing material, characterised as strip-cast leaded bronze cast onto a steel backing plate, most notably VP1, VP2 and VP10, all of which have differing amounts of lead, tin and iron elements.

Overlay Material

Overlays used to consist of lead-tin or lead-indium Babbitt metals, but as with the substrate, there is now a plethora of different overlay material options available.

MAHLE state that the overlay provides three critical characteristics – conformability, compatibility and embeddability.

Conformability – in an ideal world, the bearing housing and the journal shaft would be infinitely stiff and the alignment of the shafts would be perfectly true. Of course, in the real world that is not the case: the cylinder block and con rod assemblies will flex as load is applied, and the crankshaft journals will twist and bend, albeit by imperceptible amounts in most cases. But it means the bearing material must be able to distort elastically in response to distortions of the mating parts.

Compatibility – it is almost impossible to prevent boundary and mixed lubrication conditions during low-speed running, so there will sometimes be metal-to-metal contact between the bearing’s running surface and the crankshaft journal. The chosen material for the surface of the bearing must therefore be compatible with that for the crankshaft journal. A poor combination of materials can lead to galling or even seizure.

Embeddability – even with the best filtration systems and oil additive treatments, there will always be minute foreign particles in the oil that the bearing will have to cope with. As a result, the overlay material of the bearing must be able to absorb or embed this debris in the surface, otherwise the particles will eventually score the surfaces of the bearing and the journal at high loads.

Bearings from Modatek

The original crankshaft journal bearings that were used by Cosworth on their early engines like the DFV were developed with Vandervell. Over subsequent years these bearings have proven to be reliable in a wide range of applications including Formula 1.

We continue to use bearings from Mahle, who took over Vandervell in 2007. We now stock a wide range of Mahle bearings, including rod and mains bearings for the Cosworth YB engine.

MTK0016 YB Rod Bearings

This feature on crank bearing material is based on an article written by Modatek’s Matt Grant for Race Engine Technology, issue 114. You can purchase a back copy of this publication here: https://www.highpowermedia.com/Product/race-engine-technology-issue-114

Have you ever wanted to take a peek inside a Formula 1 engine? Up until now the insides of a Formula 1 engine have remained a guarded secret – until now! Take a look inside one of the best engines from the mid-2000s, the Cosworth TJ.

We stock race-used parts from Cosworth’s TJ and CA engines that can be displayed for all to see. Just keep an eye on our Memorabilia section of our on-line shop to see what’s currently available. Each part can be bought directly from the website shop and delivered straight to your door.

The Cosworth TJ

The TJ was the first used by Jaguar Racing in 2003. The team continued to use this engine when they morphed into Red Bull in 2005. It also saw active service powering the Jordan and Minardi teams in 2004 and 2005, respectively. Its final year of racing was in 2006, when Torro Rosso (formally Minardi) ran the engine against the V8’s.

It was a clean sheet design, learning on the lessons from the successful CK engine that had been used by Stewart Grand Prix / Jaguar Racing from 1999 to 2002. There were several important concepts copied from the CK, such as the unique central-beam cylinder head philosophy.

The TJ received an extensive amount of development, eventually allowing the engine to reach a maximum speed of 19,000 rpm and peak power in the region of 900 bhp.

Last of its Era?

When it was introduced in 2003 the TJ was the epitome of the “anything goes” regulations from that era. Devoid of rules that restricted the number of engines that could be used by each driver, the emphasis was very much of extracting the maximum out of each engine in just one session or race.

It wasn’t uncommon back then for teams to get through over six engines in one weekend (and that’s not allowing for unscheduled engine changes!). As a result, engines like the TJ were designed with performance over reliability in mind.

Midway through the life of the TJ, rules aimed at extending engine life started to creep in. In 2004 the FIA introduced the “two race” requirement that each engine had to attain. Engines like the TJ that were designed for outright power and performance had to be re-engineered to last longer.

In addition, restrictions in later years on low density materials plus minimum weights and centre-of-gravity heights meant that the lightweight features seen on the TJ became redundant in subsequent engines.

In its final year of service with Torro Rosso in 2006, the TJ’s maximum engine speed was capped at 16,700 rpm amid fears that it would outpace the newly-introduced V8 engines (including the 20,000 rpm CA from Cosworth).

However, the TJ lives on, and can be seen (and heard) running in the back of numerous historic Formula 1 cars.

Modatek are pleased to be able to help keep these engines alive for historic race series and demonstration events. If you need assistance then get in touch.

Have you ever wondered what the green coating is on the skirts of some of our Cosworth pistons? If you have, then you are not alone! The coating is actually a fluoropolymer material that has the tradename Xylan, and Cosworth have been using this coating for a number of years on a wide range of parts, and not just the piston. Oil pump and scavenge pump bodies also received the Xylan treatment to help prevent wear from rotors.

Not all Cosworth pistons have this coating – pistons from earlier engines like the BDA, for example, were designed without any skirt coating. But the increases in piston speed in their race engines meant that some form of coating was necessary to help prevent scuffing of the skirts against the cylinder bore walls.

In the early ’90s Cosworth started to coat the skirts on pistons from several different race engine categories, including those destined for Formula 1 like this one from the race-winning HB engine (the coating also looks great when etched with the driver’s name!). Cosworth soon started Xylan coating nearly all of its race and high performance engine pistons, including  the ones we sell today for the YB engine.

What is Xylan?

Cosworth PA2062 Piston Xylan Coating

Xylan was developed by DuPont in 1969, primarily for kitchenware utensils as an alternative to Teflon. Its excellent wear properties meant that it was soon adopted by the automotive industry, where it found its way into numerous different applications.

Xylan is in effect a composite material comprising of a dry film lubricant contained in a matrix with high-temperature organic polymers. This creates what can be termed as a plastic alloy that has excellent surface characteristics and is easy to apply. Xylan has a very low coefficient of friction, so its perfect for the interface between the skirt and the cylinder bore. It also provides exceptional wear resistance, and quite often Cosworth would strip engines down and see that the Xylan coating was completely unmarked. This is in part due to another benefit of Xylan – it has excellent surface adhesion.

The Cosworth Process

As with any coating, good preparation of the substrate surface is vital. In the case of pistons, the skirts need to be completely clean and free of any oil. Once cleaned, the skirt area is masked off with special tape which can simply be peeled off after coating.

The piston xylan coating is then sprayed on in one of Cosworth’s special custom-made spray booths onto the skirt. Each pass of the spray adds around 5 microns, so the thickness can be carefully built up to the required level. Most piston skirts only need a couple of passes to get the thickness up to the required level of between 6 and 12 microns.

The pistons are then left to dry – normally this can take 24 hours. Once dry, the pistons are good to go, with no other treatment necessary.

Cosworth YB 4WD 8:1 PA2062 Piston

If you’d like to know more about the Xylan coatings that Cosworth use on piston skirts then get in touch.

All of our Cosworth production pistons start life as a piston forging to ensure that they are as strong as possible.

Put simply, the forging process involves pushing a bespoke die under great pressure into the billet. As a result, the billet material flows into the required shape of the piston. The forging process gives us the finished machined shape in the undercrown. This reduces the amount of machining that is required. Further, the orientation and distortion of the grains in the material is optimised to give superior mechanical strength. Aligning the axes of the material grain in a certain way can have a significant impact on the structural properties of the piston.

Temperature Effects

Our pistons begin their life as a billet of extruded aluminium alloy bar. This billet is heated to a pre-determined temperature to ensure that the billet is soft, but not at melting point. When the billet is at the correct temperature, it is placed into the forge and shaped by the die. The forging temperature needs to be carefully controlled. This temperature will have a significant effect on the homogeneity of the microstructure. If there were localised changes in the billet temperature then this could lead to inconsistencies in the material properties of the finished piston. For example, if the die was cold then the outer surfaces of the piston would cool rapidly. This would lead to a varied grain structure in the finished piston.

Hence the die is heated to the same temperature as the billet, creating a process known as isothermal forging. This process  that keeps the billet at its maximum elevated temperature throughout the entire forging operation. During the forging process, any cooling at the interface between the die and billet is eliminated, which can greatly improve the flow characteristics and hence the grain structure of the finished item.

Cosworth YB 4WD 8:1 PA2062 Piston

If you’d like to know more about our genuine Cosworth pistons and how they can provide you with the performance and reliability that you need from your engine then please get in touch via our Contact page.

A number of Cosworth pistons have what Cosworth refer to ‘piston anti-detonation bands’, but what actually are they? The term anti-detonation is perhaps a bit of a misnomer, as these bands don’t actually stop detonation. Instead, these bands mitigates the effects of detonation.

Piston Anti-detonation Bands


Understanding Detonation

Before we take a closer look at these bands, what actually is detonation? It’s a phenomena that is commonly mentioned in the world of high performance engines, but it does sometimes get used incorrectly. Detonation is broadly defined as unwanted or abnormal combustion. It occurs when both the temperature and the pressure in the unburnt mixture of air and fuel exceeds a critical level.

In a normal combustion event the spark plug will ignite the air/fuel mixture inside the combustion chamber. This happens before TDC whilst the piston is travelling upwards. The ignition of the air/fuel mixture creates a flame front that spreads outwards, igniting more of the mixture and leaving behind the burnt gases.

However, in certain circumstances when the temperature is hot enough, the unburnt mixture that is furthest from the spark plug (towards the wall of the cylinder) will ignite before the flame front reaches it. This unwanted and uncontrolled combustion can happen almost instantaneously, setting up strong pressure waves that hit the walls of the cylinder. These shock waves create the distinctive pinging noise that can be heard and indicate the presence of detonation.

Cosworth Anti-detonation Bands

In the early ‘90s, one step that Cosworth took to reduce the effects of detonation was to introduce anti-detonation bands. These grooves are located above on the top land above the top ring groove. They are typically around 0.2 mm deep and 1.25 mm wide. Most pistons will have two, three or four of these grooves, depending on the height of the top land.

But these shock waves can also be extremely destructive – they can inflict damage to the piston crown and in particular to the area around the top land. If detonation is severe then ultimately the piston will be irrevocably damaged, with the potential of engine seizure.

There are a number of Cosworth pistons that have anti-detonation bands, ranging from those for the YB through to some of the more recent Formula 1 engines. Basically, any engine that will experience high cylinder pressures will benefit from the usage of anti-detonation bands.

Cosworth YB Piston


The aim of these bands is to protect the top ring by disrupting any high-pressure pulse waves from detonation. The volume within these grooves, although small, allows a space for additional atomisation of the fuel and air mixture. When detonation occurs, this extra space provides an outlet for the shock waves.

These bands also bring about a couple of additional benefits. Firstly, they can prevent the build-up of carbon above the top ring which would otherwise cause the ring to stick in the ring groove. Secondly, they reduce the amount of contact between the top land the cylinder bore wall. This makes them very beneficial at high engine speeds and piston temperatures. Hence they are also occasionally referred to as “contact reduction grooves”.

The introduction of anti-detonation bands by Cosworth was another example of their experience and understanding of internal combustion engines. Indeed, when owned by Vickers, Cosworth actually patented their anti-detonation band design in the US, take a look here : https://patents.justia.com/patent/5267505

Anti-detonation Band

Want to know more about the range of  Cosworth pistons that we supply? Then get in touch via our Contact Us page.

One of the most important aspects of any IC engine is the correct timing of the rotational movement of the camshafts relative to the reciprocating movement of the pistons. Any errors in this timing will result in detrimental performance, and could in the worst case lead to contact between the valves and pistons, which could prove to be both catastrophic and costly. There are a number of ways that this meticulously choreographed movement of valves can be achieved, such as with gears or chains. But one of the most common methods in a road car is with the use of a timing (or cam drive) belt. The timing belt is normally driven by the crankshaft, and then turns pulleys that drive the camshafts.

Cosworth 20019488 YB Heavy Duty Timing Belt


Timing Belt Construction

The majority of automotive timing belts are constructed from an elastomer body that contains tension cords, with a fabric backing and a tooth jacket.

Looking at these four ingredients one by one, then the first is the elastomer body. This is normally a high temperature spec rubber such as HNBR (hydrogenated nitrile butadiene rubber) or EPD (ethylene propylene diene). HNBR is more suited to engines, as it can handle exposure to lubricating oils. The elastomer body can sometimes be reinforced with aramid fibres, which help strengthen the belt and provide extra protection for the teeth. Aramid is a heat-resistant and strong synthetic fibre that is sometimes referred to by the tradename Kevlar. Our heavy duty belts for the YB engine contain this aramid reinforcement, which helps make them three times stronger than a conventional automotive timing belt.

The tension cords help to give the belt incredible levels of tensile strength without compromising on flexibility. Tension cords are typically manufactured from high-strength glass fibre, and the individual strands of fibre are bundled together and twisted for added strength. It is the presence of glass fibre that means that a timing belt should never be ‘crimped’, which is the action of over-bending or twisting the belt that then shears the glass fibres.

Next up we have the tooth jacket, which is a temperature-resistant polymide fabric that helps to protect the teeth from abrasion as well as shear forces. The last ingredient is the fabric backing, which is usually another type of polymide fabric. This fabric backing is used on the smooth face of the belt that will run against the belt tensioner, so resistance to abrasion and wear is a must for the fabric.

Cosworth 20019488 Heavy Duty Timing Belt

Toothed Belts

In the majority of cases the timing belt is driven by teeth on the belt that engage with matching teeth in the pulleys on the crankshaft and camshafts. It is these teeth that provide the accuracy in timing, and hence toothed belt drives are often terms ‘synchronous belts’, as the keep the movement of the cams in synch with the crankshaft.

Keeping the teeth engaged at all times is vital. If a tooth jumps out of position then this is called ‘ratcheting’ (some people might use stronger language when this happens!). It is normal for the belt teeth to try to escape from the adjacent teeth in the pulley – when this happens the belt tension increases, which pushes the teeth back together, but this can lead to long lasting damage of the tension cords inside the belt. Complete slippage of the belt teeth out of the pulley teeth usually happens because either the tooth engagement is poor, or because of lack of belt tension.

The profile of the teeth is an important factor for the performance of the timing belt, and there are generally three categories for the profile – trapezoidal, curvilinear or modified curvilinear. Curvilinear profiled belts are sometimes called High Torque Drive (HTD), and modified curvilinear profiled belts are often called Super Torque Drive (STD), S-type Tooth Profile Dual-sided (STPD) or GT.

Belt Tooth Profiles

Tooth Profiles

The trapezoidal profile is the oldest of the three, introduced over 80 years ago. The shape of the profile is a trapezium, with straight flanks that are angled inwards towards the tip of the tooth. When the teeth revolve around the pulley, these flanks create an involute curve that matches the involute tooth profile on the pulley.

One disadvantage with the trapezoidal tool profile is that it has sharp corners at the root of the tooth, and this can create high stress concentrations that can weaken the belt. To overcome this, the curvilinear has fully radiused corners, which helps to even out the stresses. Also, the curvilinear tooth profile is taller than the trapezoidal profile, making it more difficult for the teeth to jump out of position and also giving a larger contact area, which in turn helps to reduce both stress and noise.

The last of these three profiles, modified curvilinear, is today the most popular type with timing belt manufacturers. As the name suggests, this profile is based on the curvilinear profile, but has a shallower tooth height along with an increase in flank angle. These changes help to give the modified curvilinear profile the ability to withstand higher amounts of torque.

Cosworth Engine Belts

Want more information on the belts that we stock for Cosworth engines? Then get in touch via our contacts page.

Our range of Cosworth YB pistons feature an offset gudgeon pin, which helps to reduce piston skirt wear and engine noise. To understand why, we have to consider the loads that the piston will experience.

Cosworth YB Piston Drawing

When the piston moves up and down the bore the small end of the connecting rod will articulate forwards and backwards around the pin. This creates a loading on the piston that pushes the piston sideways. This side load varies with crank angle and is also different for each stage of the four-stroke cycle. The largest side load occurs during the power stroke, when there is a combination of inertia and gas loads that will push the piston sideways towards what is referred to as the major thrust side.

Side Loads

These side loads can have an extremely detrimental effect on the operation of the piston. For example, they can promote tilting or rocking of the piston in the bore, which in turn can increase wear at the top of the lands and at the bottom of the skirt. This type of movement can also create vibrations that are then radiated through the engine and can be heard outside the engine.

One way to reduce the major thrust side load is to offset the pin’s centre away from the cylinder bore centreline and towards the major thrust side. Even just a small offset can have a noticeable effect in reducing wear and noise during running.

If you’d like to know more about the science behind piston offset pin tech, or about the YB pistons that Cosworth makes exclusively for Modatek, then please get in touch via our Contact page.