A fundamental limitation of the inverted bucket tappet is the opening and closing velocity of the valve. The valve spacing dictates the spacing of the tappet axes and the maximum diameter of the tappet. This maximum diameter in turn limits the valve opening velocity. If we assume that the follower’s axis intersects the camshaft axis (most do) then it can be proved that the instantaneous distance from the follower axis to the cam-to-follower contact is equal to the lift velocity, when the velocity is measured in mm (or any other unit of length) per radian of cam rotation.

If we had a 30 mm diameter tappet, the maximum velocity - assuming we were happy to sweep the cam profile contact point right to the edge of the tappet - would be 15 mm per radian. In practice, the limit is lower than this, owing to edge detailing on the top of the tappet and a desire to keep a certain finite width of lobe sitting flat on the face of the tappet. If we wish to use a symmetrical lift profile, it is unlikely that we will want there to be an offset between the camshaft axis and the follower axis, lest we limit velocities further.

If the valvetrain designer cannot increase the diameter of the follower then he can avail himself of extra velocity by using a tappet with a curved top. This has the same geometric effect as changing from a flat-faced lifter to a roller-lifter on a pushrod engine. While maintaining the same valve lift profile, the point of contact remains closer to the follower axis for any value of pressure angle and follower curvature. For a given pressure angle, a smaller cam follower curvature decreases the distance from the follower axis to the point of contact. This can be very handy if your new valve lift profile developed with your engine simulation software overhangs the edge of a flat tappet and you are prevented by engine geometry from increasing the tappet size.

Followers with curved contact surfaces have disadvantages though. They are taller and heavier than their flat counterparts, and the increase in mass can require a different valve spring. Such followers also require there to be an anti-rotation feature on the follower which locates with a corresponding feature in the follower bore. In order to maintain the correct alignment of the follower with the cam lobe, we therefore need to increase the complexity and cost of the follower and the machining in the cylinder head or cam carrier.

Written by Wayne Ward

It is often thought that this sort of analysis is the preserve of manufacturers building clean-sheet engine designs, but the benefits of combustion analysis make it relevant even to small-scale race engine builders and tuners. The number of parameters that can be investigated given the equipment available on the open market is legion, although for the purposes of this article we will look at just one of the most useful, the pressure-volume (PV) loop. The reason is twofold - it provides a very clear graphical representation of a cylinder’s behaviour, and any changes or corrections made to an engine are immediately visible.

Equipment
The production of a PV loop for an engine requires at least one pressure sensor per cylinder, a means of accurately recording crank position, and computing resources to gather and present the recorded data. For purely experimental engines, pressure sensors can be built into the cylinder head architecture, but a far more practical approach is the use of spark plug sensors. These are specially designed plugs that incorporate a high-speed piezoelectric pressure sensor that can record the pressures generated in the cylinder.

There are several companies producing such sensors, which come with features such as ‘flame guards’ to reduce errors produced by thermal shock experienced during the combustion process. Crank position can be measured using an engine’s built-in sensors, but these generally lack resolution so specialised sensors are preferable. Most companies specialising in such test equipment will normally be able to supply an integrated data acquisition system complete with plug-and-play interfaces for sensors and dedicated software to process the large volume of data generated.

The PV loop
A PV loop is a line graph that displays in-cylinder pressure at all points in a particular cylinder’s cycle. This simple graphical representation provides an invaluable insight into the combustion behaviour of an engine as the cylinder volume changes with the crank stroke. This representation also allows for very easy comparison of different engine configurations.

Fig. 1 - The PV loop representing the ideal Otto Cycle

Fig. 2 - A typical PV loop for a spark-ignition engine

Fig. 1 shows the ideal PV diagram for an Otto cycle engine, while Fig. 2 shows the PV diagram for a typical combustion engine, with the key events annotated. It can be seen from the comparison of the two diagrams that the ideal cycle has no pumping loop (grey shading). This indicates that the gas exchanges from the intake manifold into the cylinder, and from the cylinder into the exhaust manifold after combustion, should ideally occur without any associated losses. In practice, this can never be realised, and work is always expended drawing air into the cylinder and expelling exhaust gas from the cylinder.

The work output of an internal combustion engine is indicated by the difference in area contained within the power loop and the pumping loop. This means that improvements in an engine’s efficiency can be identified by an increase in the size of the power loop relative to the pumping loop. In simple terms, if the combustion efficiency of the engine is increased, this will be represented as a greater area within the power loop, with no change in the pumping loop.

However, useable power can also be gained by reducing pumping losses, which can be identified by a reduction in the area of the pumping loop while the power loop remains constant. With an engine instrumented to record this information, the effects of changes to factors such as valve events, ignition timing and inlet/exhaust design can be compared very easily.

Producing PV graphs for individual cylinders also allows each cylinder to be compared with the others, and treated as an individual engine, the goal being to make all cylinders as efficient as possible. This is the key benefit of combustion analysis: where a dyno allows you to look at only the overall output of an engine, individual cylinder analysis gives a far more complete picture of an engine’s performance.

In future instalments of RET-Monitor we will look at the individual areas of the PV graph in more detail and relate them to potential performance gains, as well as address some of the other aspects of combustion analysis.

Fig. 3 - The pumping section of PV graph for a six-cylinder engine. Each line represents the pressure trace for a separate cylinder. Note the variation between the values for each cylinder

Written by Lawrence Butcher

Where there is sliding involved in a loaded non-conformal contact, it is common to find parts with a very fine surface finish. This is not simply for decoration; there are very good reasons for achieving a good finish, and the cam-to-follower contact is an excellent example here. While gears are a non-conformal contact pair, there is a lot of rolling contact in a gear, while the cam-to-follower contact is predominantly one of sliding.

The lubrication of these parts is very important to ensure sufficient service life and to mitigate the effects of friction. The lubrication regime is fundamentally hydrodynamic, and relies on oil being swept into the contact. There are two points during the opening and closing cycle of the valve where the entrainment velocity - a measure of the speed at which oil is swept into the contact - becomes zero. Where entrainment velocities are zero, or close to it, the oil film will tend to become much thinner. We can sometimes see the point at which the entrainment velocity is low by examining the cam lobe, as it can show signs of distress before any other points. If you go through the calculations, which are beyond the scope of this article, you can often correlate this point against real cam lobe damage.

As the oil film becomes very thin, the height of the asperities (high points) on the surfaces of the contacting components becomes significant. At the point where these can begin to touch because the oil film isn’t thick enough to keep them apart, friction starts to increase rapidly. In this situation, the lubrication regime is called ‘mixed’ or ‘boundary’ lubrication, where the applicable friction coefficient is a function of both the lubricated friction coefficient and the dynamic dry coefficient of friction. As the film thickness ratio - the ratio of oil film thickness to the ‘combined’ surface finish of the pairs of surfaces in contact - decreases, a greater proportion of the contact area comes into solid contact.

By improving surface finish, the oil film can be much thinner before boundary lubrication applies. This mitigates frictional losses and can prevent wear from occurring.

There are several possible wear mechanisms in such a contact. Even if adhesive or abrasive wear does not apply, subsurface fatigue may become a problem, and another effect of rising friction coefficient is to increase the level of subsurface stress.

This is why it is common to find very highly polished camshafts and followers.

Fig. 1 - Cams are often polished, and there are very good reasons for doing so

Written by Wayne Ward

There are reasons why designers of new engines might wish to have the cylinder axis not intersect the crankshaft axis, and these have been used by production vehicle manufacturers in justifying and producing new engines with non-conventional layouts for many years. The two reasons for producing an engine with this layout have opposing results. The first, which does not apply to race engine design, is to improve the noise characteristics of the engine, and specifically to reduce the noise due to piston slap at top dead centre (TDC).

The second reason, which is definitely of more interest to motorsport engineers, is to improve engine output by reducing friction. By having the cylinder axis offset from the crankshaft axis, the aim is to keep rod angularity to a minimum while the cylinder pressure is at a maximum. The reasoning behind the concept is that reduced angularity leads to lower piston thrust forces, and therefore lower frictional losses during the period of maximum cylinder pressure. In practice, the concept is much more easily applied to inline engines than to vee engines. For those whose business it is to develop production-based engines, the concept is difficult, if not impossible, to apply retrospectively.

However, there is an equivalent mechanism that does allow the concept to be applied, and that is to have the pin bore offset in the piston. This allows the same crank/rod pin axis geometry as an engine designed with offset cylinder axes, but within a ‘conventional’ engines. For those who have pistons custom-made, the option is open to design such a piston, but for some popular engines, pistons with offset pin bores can be bought from motorsport piston manufacturers.

The amount by which the cylinder axes need to be offset depends on many variables, but the main ones are the ratio of crank throw to con rod length and the angle after TDC at which maximum cylinder pressure occurs. The design engineer may have the simple aim to have the con rod parallel with the cylinder axis at maximum pressure, or he may resort to more complex simulation in order to reduce friction over a given range of engine speed.

Depending on the angle after TDC at which maximum cylinder pressure occurs, this may preclude being able to practically apply the desired pin bore offset within the piston. In this case, the engineer will have to settle for the pragmatic approach and be satisfied with whatever gain is observed. The practical limit to moving the pin bore off centre is often the excessive moment due to the large distance between the pin bore axis and the centre of gravity of the piston.

Another point to be aware of is the effect on the engine stroke due to offsetting the pin bore. For a given crankshaft stroke, offsetting the pin bore or cylinder axes increases the stroke, and this may be enough to put the engine beyond the capacity limits of the class in which it is being raced.

Fig. 1 - This production engine piston uses pin bore offsetting

Written by Wayne Ward

The matter of fluid movement within enclosed spaces has been the subject of a great deal of study in a number of areas outside motorsport, notably in the marine industry and among mainstream automobile manufacturers. The first efforts at reliably simulating the behaviour of oil in tanks was in fact undertaken to understand its movement within bulk crude carriers in rough seas - in essence giant oil tanks!

Advances in modern CFD/FEA capability and high-performance computing systems mean it is practical now for racecar designers to incorporate slosh simulation (encompassing any fluid tanks, not simply oil) into the overall design process. This reduces the need for physical testing and reduces the chances of component failure on track due to oil starvation, while also allowing the behaviour of fluids to be taken into account in relation to a car’s dynamic behaviour.

The interaction between a fluid in a tank, the tank and any air within the tank is a very complex one, and it is only recently that commercially available CFD/FEA packages have been capable of simulating such interactions. The two key methods used to simulate conditions are VOF (volume of flow) and arbitrary Lagrangian-Eulerian (ALE) formulations, both of which are able to account for the interactions between material and fluid as well as the boundary flow conditions on the surface of the fluid.

The VOF method is a numerical technique for tracking and locating the fluid-fluid interface. It belongs to the class of Eulerian methods that are characterised by a mesh that is either stationary or is moving in a certain prescribed manner to accommodate the evolving shape of the interface. As such, VOF is an advection scheme - a numerical recipe that allows the programmer to track the shape and position of the interface, but it is not a standalone flow-solving algorithm; the Navier-Stokes equations describing the motion of the flow have to be solved separately.

Several commercial CFD packages integrate VOF capability, although it is only in the past few years that the ability to conduct true VOF simulation has been incorporated into them. VOF methods are very effective for calculating the behaviour of fluids but are limited to this role, so if other factors such as heat transfer from the fluid to container material need to be studied, a different set of codes is needed. This is of particular note in racecar design, where items such as the oil tank can be packaged very closely to other components, where heat transfer could be an issue.

It is here that newly developed codes based around an ALE method can prove very beneficial. The ALE is a finite element formulation in which the computational system is not fixed in space as with the Eulerian-based VOF, or attached to material as in Lagrangian-based finite element formulations. When using the ALE technique in engineering simulations, the computational mesh inside the domains can move arbitrarily to optimise the shapes of elements, while the mesh on the boundaries and interfaces of the domains can move along with materials to precisely track the boundaries and interfaces of a multi-material system.

While very effective for fluid movement simulation, the big advantage of ALE-based finite element formulations is their ability to be reduced to either Lagrangian-based finite element formulations by equating mesh motion to material motion, or Eulerian-based finite element formulations by fixing the mesh in space. This means that only one FEA code is needed to perform the range of simulations from fluid flow and fluid-structure interactions through to heat transfer.

Such capability is not widely available in commercially produced packages, but this is likely to change soon. Whichever approach is chosen, it is now well within the abilities of teams with a good computing capability to accurately assess the behaviour of fluids throughout the vehicle and optimise designs to ensure that the performance or reliability of systems is not compromised.

Fig. 1 - Fluid slosh simulated in OpenFoam, a feely available open source simulation package

Written by Lawrence Butcher

Consisting of two internal rotors in the shape of lobes rotating together, and phased using gears to prevent contact, when geared to the crankshaft the air delivery per revolution is fixed, so as the engine speed increases so does the air delivery of the pump. Unlike the turbocharger, no fancy pop-off valves or bypass systems to avoid excessive boost pressures have to be incorporated. And since air delivery is broadly linear with the engine speed, the manifold or boost pressure will remain more or less unchanged.

Not fully appreciated by many, supercharged engines simply add to the driving experience. Opening the throttle produces instant power and the feeling of driving a much larger engine than the one under the bonnet, and because of this (and unlike turbocharged machines) driveability - that characteristic that often gets lost in the search for power - is vastly improved; ask anyone who has conducted back-to-back tests with a turbocharged vehicle. With a supercharged engine you quickly forget (if you ever knew) the presence of the supercharger, unlike a turbocharged unit where the delay in response, however slight, is always there, even with the best of systems.

Strictly speaking, a Roots-type supercharger is a pump rather than an air compressor, and while more modern mechanical units may include a small amount of internal compression, generally speaking these machines have no internal compression inside them; the increased intake plenum or boost pressure is created solely by the restriction to flow in the engine.

When running at a part-load, off-throttle condition - which let’s face it, even competition vehicles do for quite a lot of the time - there is therefore little parasitic drag, unlike say that of a turbine of a turbocharger in the exhaust stream, which will always impose some level of back-pressure against the engine. On some newer units these twin lobes have been replaced by a three-lobe design, which when twisted to form a helix along the length of the rotor introduce an element of compression to the intake charge. More important for road-based applications, this reduces the intake port pulsation and therefore intake noise.

Since Roots-type machines have no contacting parts, friction within the device is low, and the design of the intermeshing rotors is such that no out-of-balance forces are generated. Roots blowers are therefore safe up to quite high speeds, with 4000 rpm being quoted by some manufacturers. At these speeds, however, designers need to take heed of the increased inertia of the rotor and design the drive system to cope.

One particular issue with Roots superchargers is the clearance between the rotors. Machined and coated to minimise this at all times but never to make contact, at low speeds pumping efficiency can be impaired and lead to a certain amount of leakage. In modern designs this has been all but eliminated, and reliability has been vastly improved such that when specified on OE vehicles, units can be expected to last the vehicle’s life.

So while modern trends seem to be leaning towards complex turbocharger systems, for raw performance and simplicity of installation - as well the sheer driving delight - you simply can’t beat a good blower.

Fig. 1 - A Roots-type supercharger

Written by John Coxon

Complicated by the phasing out of four-star 97 RON fuel and replacing it with a 95 RON - so-called ‘premium’ unleaded - the fuel companies produced a 98 RON lead replacement fuel (LRP) that replaced the octane and trusted enthusiasts to fit hardened valve seats to contain the wear. And when sales of the more expensive LRP fuel decayed a few years later, the compromise fuel was quietly dropped from the forecourt. By then specialist fuels were on the market in small amounts for owners who absolutely needed them, and those remaining used approved lead replacement additives.

These days, however, the issue is one of ethanol, and while the lead issue was all about the IQ levels of children living close to urban motorways, now the issue is one of sustainability and environmental considerations - and a UK government initiative called the Renewable Transport Fuel Obligation. The current standard for UK road fuel, BS EN 228, allows for up to a maximum of 5% of ethanol by volume. In practice, according to the individual refinery and the issues surrounding the transportation of the hygroscopic (water-absorbing) ethanol, not all fuels will contain this level, and fuel sold will contain either no ethanol or about 4-5 % depending on where you live. At this level there is no requirement to identify the ethanol content, and in most cases the general level of the fuel’s performance is unchanged in any case.

However, with the advent of the RTFO by some now unspecified time in the future (at one time rumoured to be 2013 but now refuted by the Department for Transport and Rural Affairs) this maximum content is to increase to 10% by volume. And if 5% didn’t exhibit any truly discernable difference then 10% most probably will. For modern vehicles with fuel injection and adaptive control designed for modern fuels there are no issues, but for those with carburettors (remember them?) many pre-2000 vehicles and even early direct-injection designs, the problems could be significant.

These issues fall into three main categories - combustion, corrosion and compatibility.

Compared to the value of 14.7:1 for non-ethanol, ‘normal’ fuels, E10 (as it is now called) has a stoichiometric value of 14.1:1. If the specific energies are broadly the same while the performance given by each will be similar, in those vehicles using carburettors, engines will run about 4-5% lean. For fuel economy this might be acceptable, but if mixture strengths are not adjusted to compensate, valve seat burning could result. The greater volatility of the increased amount of ethanol could also introduce fuel vapour lock in the under-bonnet fuel lines, so these will also have to be routed away from sources of heat.

It has long been appreciated that ethanol-gasoline blends encourage corrosion when exposed to aluminium. The precise mechanisms involved, however, are complicated. Ethanol would seem to be a better conductor of electricity than hydrocarbon fuels, and the galvanic effect of dissimilar metals in the presence of absorbed water may have much to do with it. For racers the answer must always be to empty the fuel tank and run the engine on neat gasoline at the end of the day, but where this is impractical, galvanic action as a result of the proximity of dissimilar metals (such as aluminium-steel, aluminium-brass or zinc-brass) has to be avoided.

Elastomer compatibility also has to be taken very seriously when using the different fuel blends. Seal swell characteristics or the progressive hardening of elastomers can lead to system leaks, and it appears that, as in the case of corrosion, ethanol-gasoline blends are even more aggressive than either pure gasoline or pure ethanol alone. In general, however, the later types of fluorinated seal compounds are more resistant to attack than non-fluorinated types, and the greater the degree of fluorine then the higher the resistance.

In modern fuel systems these compatibility issues will all have been sorted, but being ‘green’ and protecting your fuel system for those with historic or classic vehicles may not be that simple.

Fig. 1 - Comparison of fuel characteristics

Written by John Coxon

The forces the shaft is subjected to are substantial, so for many years the only option was to use metallic materials - usually steel or in some cases aluminium - for their construction. Despite providing the strength required to achieve the durability needed with high-output race engines, especially when used for endurance racing applications, metallic prop shafts have some disadvantages. Notably they are heavy, which adds to vehicle weight and increases parasitic power losses, but more important the material choice severely limits the maximum shaft length used in high-rpm applications.

This is due to the bending resonance experienced by a tube as it reaches its critical speed. This resonance is described by the following equation:

where:
Nc is the critical shaft speed
L is the shaft length
I is the tube’s second moment of area
A is the cross-sectional area of tube
C is the constant vibration
E/ñ is the specific modulus of the shaft material

The resonance will destroy a prop shaft, and while that’s not a huge problem on low-revving mainstream automobiles, it is a serious issue in high-speed race applications. The only ways to increase the critical speed are to increase the diameter of the tubing or increase the specific modulus of the material. Space constraints invariably set a limit the diameter of the tube, while it is an interesting characteristic of metals that the specific modulus is approximately constant, despite big differences in density.

One driveshaft manufacturer discovered that at a speed of 8000 rpm the longest steel shaft that could be used before resonance became a problem was only 1250 mm. This is too short for most applications, so the general approach is to use a two-piece shaft with a supported joint in the middle. Evidently from a racing perspective, that is not ideal, as it introduces both extra weight and complexity into the system.

The solution to this problem was the introduction of composite prop shafts, the use of which has been made possible by advances in materials technology. By using fibre-reinforced composites, it is possible to orientate the fibres in a tube’s structure so that the bending modulus has a high value (above 100 GPa) while the specific gravity is low (below 1.6). This leads to a favourable specific bending modulus and enhanced critical speed.

This is not a particularly new idea, and the first composite shafts began to appear in production cars in the late 1980s, although these used an aluminium core reinforced with a carbon fibre outer sleeve. Composite prop shafts were experimented with in the 1990s in rallying and circuit racing, but there were many failures. However, the past decade has seen great improvements in both composite materials technology and, more important, the ability to simulate and assess different construction techniques. The result is that nearly all competition GT cars now feature composite shafts, which have proved to be more than up to the required tasks.

Making these shafts is a complex process, and predominantly uses a filament-wound carbon tube for the main structure. Manufacture involves winding filaments under varying amounts of tension over a male mould, or mandrel. The mandrel rotates while a carriage moves horizontally, laying down fibres in the desired pattern. The exact pattern of the ‘lay-up’ will be determined using FEA software and specialist composite CAE packages, to ensure the fibres are oriented in the optimum direction to absorb loadings. The connections at the tube’s ends are usually a very tight interference fit, and bonded to the tube with high-strength resin.

The overall result is impressive. One manufacturer says a representative 1.5 m prop shaft for GT use weighs only 2 kg, compared to about 10 kg for a similar two-piece steel item. Yet despite the low weight the shaft can still withstand up to 3500 Nm of torque while operating in an environment at 100 C. This represents a major improvement in both overall vehicle weight and rotating mass, and makes it easy to see why nearly all front-engined, rear-wheel-drive GT cars now run composite props.

Fig. 1 - Composite prop shaft for use in endurance racing

Written by Lawrence Butcher

Given that we know the pitch of the bolt, could we not simply arrive at an angle through which we turn the fastener (nut or bolt)? If we assumed the joint (clamped) members to be infinitely stiff, we could simply work back from our knowledge of the axial bolt stiffness and pitch to correlate tightening angle and force.

If the fastener has a stiffness k and pitch p, then for a single complete turn of the fastener head, we assume that the fastener has stretched by a distance p, and that the load is kp. For example, if the fastener stiffness is 50,000 N/mm and the pitch is 1.25 mm, then a single turn should increase the load by 62,500 N, if we assume that the joint is infinitely stiff.

Of course, we have immediately raised the first difficulty - the joint members are certainly not infinitely stiff. The finite stiffness of the clamped members must be taken into account, and if we imagine that the clamped parts are made of rubber, several turns of the fastener described above will develop very little load. Unless the geometry of the joint materials is extremely simple, the calculation of their stiffness is not a trivial matter, and the options for engineers without access to finite element analysis (FEA) software packages are either to turn to physical testing or to calculate the stiffness of the joint members based on assumptions and physical measurements.

The calculation of load is made more complicated by the finite stiffness of the clamped joint members, but there are other factors at play too. If the nut is not in contact with the washer when the nut is turned, no tensile load will be developed in the fastener, even if there is a resisting torque (for example from a locking mechanism). That much is obvious, so we may choose not to start measuring the angle of turn until the nut is in contact with the washer.

However, even with the knowledge of accurate fastener and joint stiffness, that is not going to give us the correct relationship between angle of turn and pre-load. The reason is that there is a portion of the tightening event during which the behaviour of the joint is non-linear. Until a certain load is reached, the gradient of the load/angle curve is not constant, and increases until it is equal to the gradient of the linear portion of the curve. If we assume that the load/angle relationship is linear, we stand to have significant differences between the actual fastener load and that desired, especially in cases where the design load is low and the non-linear portion of the chart takes up a significant part of the design load.

Fig. 1 - Knowledge of fastener stiffness, thread pitch and tightening angle won’t guarantee accurate bolt pre-load

Written by Wayne Ward

The Honda in particular has strayed furthest from what everyone perhaps expected these engines to be. It has the airbox and inlet on the front of the engine and the exhaust port pointing backwards in the bike, a layout not commonly used for motorcycles. One might expect that the exhaust pipe would exit straight towards the rear of the motorcycle, coming out underneath and at the rear of the seat unit. However, Honda sweeps the exhaust forwards in the motorcycle, and then down, allowing a silencer to sit just behind and below the engine.

By comparison, the KTM and Ioda engines are much more akin to a conventional motorcycle layout, with the inlet on the rear of the engine and the exhaust port on the front. As per conventional practice, the KTM and Ioda engines bring the exhaust down the front of the engine. We will have to wait to see what GE chooses to do here.

Why does the Honda bring the exhaust around the front of the engine, having gone to the effort of designing an engine with the exhaust port on the rear of the cylinder head? Well, it is probably because it is an easy way to package the necessary exhaust length. The straight-run approach from the exhaust port to the back of the bike may not have given enough length for the exhaust to tune properly without having to design a contorted pipe. If the engines require a silencer, then having this mass low on the bike is probably another reason not to use an underseat exhaust exit.

It seems that the exhaust packaging could be more easily achieved if the cylinder orientation was conventional. If we consider the exhaust alone, this is probably true. However, the small cylinder requires only a small airbox, and this can easily be housed above and ahead of the engine. Giving the inlet a ’straight run’ into the airbox will help maximise pressure at the inlet port, which is good for engine performance. It is probably the case that the design of the exhaust and the orientation of the cylinder are dictated by considerations of engine breathing.

Fig. 1 - The KTM Moto3 engine has a conventional cylinder and exhaust layout

Written by Wayne Ward