Archive for category: Technical

At the heart of an engine’s lubrication system is the oil pump, which has to continuously move oil around the complex maze of passageways inside the engine. A correctly working oil pump is vital for engine health – just one instantaneous drop in oil pressure could be catastrophic for the lubrication of bearings. But how do oil pumps work, and what are the different types of pumps that can be used on race engines?

The pump moves the lubrication fluid by trapping the fluid between moving elements. Irrespective of the type of pumping elements used, these pumps are commonly referred to as positive-displacement pumps. In a positive-displacement pump, the internal elements move to create an open void that expands and fills with the fluid. The elements then move, taking the fluid with them, after which the elements compress the volume of fluid, forcing it out of the pump.

Positive-displacement pumps can be broken down into two groups: rotating and reciprocating. The vast majority of oil pumps we find on race engines tend to be of the rotating variety, as the rotating action of the elements gives a relatively smooth flow of oil and can be designed to be compact and light. Rotary drive can be provided directly from the crankshaft via a gear, belt or chain, making them easy to package into either a new clean-sheet engine design or onto an existing engine’s architecture.

Internal Gear Pumps

The anatomy of an internal gear pump is relatively simple – an inner gear with external teeth runs inside an outer gear with internal teeth. The outer gear is housed in a cylindrical body with inlet and outlet ports on the end. The inner gear will typically have one tooth fewer than the outer gear, and the two gears run on fixed eccentric axis.  When the two gears rotate, the volume created between the meshing teeth will expand and contract to pump the fluid.

Internal gear pumps include a crescent shape in the void under the inner gear. This crescent will divide the two rotors and act as a seal between the high- and low-pressure areas.

Gerotor Pumps

A sub-set of internal gear pumps is the gerotor pump, short for generated rotor. The gerotor design does away with the crescent, creating a compact and simple solution, like our Cosworth YB oil pump.

With only two moving parts, gerotor pumps are a simple and versatile product, and it’s easy to see why they have been so popular over the years in road as well as race engines.  Gerotor pumps are especially common as oil pressure pumps, as they can have a lower fluctuation in output pressure when compared with an external gear pump.

Another plus point for the gerotor design is that the housing needs only one cylindrical bore machining in it to accommodate the outer gear, unlike an external gear or lobe pump which needs two rotor bores. And the packaging of one gear inside another means the overall size of a gerotor pump can be quite small compared with other pump options.

One other advantage that the gerotor pump has over other types is thanks to the design of the rotor teeth. These can have extremely tight clearances, which leads to less leakage between the teeth and hence better efficiency.

Note though that the tight clearances can also be detrimental, as the rotors are more prone to damage from debris. That makes gerotor pumps less attractive for scavenge pumps, where debris in the oil scavenged from crankcase is more prevalent.

Surface wear on the teeth of gerotor pumps can also be less than that seen on those of external gear pumps. That is because the relative velocity between the inner and outer gears is quite low. Take for example a gerotor pump with five teeth in the outer rotor and four in the inner rotor. At a pump speed of 2,000 rpm the relative rotational speed between the inner and outer rotors will be only 400 rpm.

Trochoid Gear Profiles

The devil is of course in the detail, and in the design of the gear profiles this maxim is especially true. Most gear profiles in internal gear pumps are from the family of trochoid curves, which is the path created by a fixed point on a circle rolling along a straight line.  

A dissection of the mathematical formulae used to generate a member of the trochoid curve group is beyond the scope of this blog, but we can touch on the definition of some of the types of trochoid curves and look at a few of the different gear profiles used by manufacturers for their ranges of gerotor pumps.

The term ‘trochoid’ actually covers a number of different subsets of curves, including epitrochoids and hypotrochoids. An epitrochoid curve is formed by the movement of a fixed point on one circle rolling around the outside of another circle. By contrast, a hypotrochoid curve is created by rolling one circle around the inside of another.

If you’ve ever played with a Spirograph toy then you will remember the motion of a pen pushed through a hole in the small cog as it rotates inside or outside a larger toothed disc. These curves are actually epitrochoids and hypotrochoids.

In a standard trochoid profile gear set, the inner gear is epitrochoidal but the outer gear isn’t and hence the profiles of the inner and outer gear teeth don’t mesh perfectly. As a result, standard trochoid profile gear sets can tend to have lower sealing properties (some of the oil from the high-pressure side will escape back to the low-pressure side) when compared with other gear profiles, giving rise to a lower volumetric efficiency.

A true epitrochoidal gear set has epitrochoidal profiles on both the inner and outer gears, and the profiles will fit perfectly inside one another when they mesh. This variant will provide a good compromise between volumetric efficiency and resistance to wear. Conversely, this style of profile can be the most prone to damage from debris, as there are no gaps for the foreign material to reside in when it passes through the teeth.

Hypertrochoid curves aren’t normally used to construct the gear profiles on automotive internal gear pumps. That is mainly because such profiles can create abrupt intersections between the lobes, which can lead to premature wear. The sharp faces of the inner gear can quickly damage the lobes of the outer once they are subject to the high speeds seen in a race engine oil pump.

One other profile used in gerotor pumps is termed the K-floid, which is essentially a hypotrochoid curve and an epitrochoid curve stitched together. K-floid gear profiles are reputed to offer good resistance to debris, but they can end up with low volumetric efficiency and are more vulnerable to wear when compared with other profiles.

The most prevalent gear profiles on the market are generated from trochoid curves with epitrochoidal inner gears or matched epitrochoidal inner and outer gears.

Some gerotor pumps also feature an asymmetric tooth profile on the outer gear. This can make then more efficient, but the downside is that the root radius tends to be smaller, leading to large stress concentrations, making them less suitable for high-performance applications.

Tooth Factors

Aside from the tooth profile, there are other changes in geometry that can be altered to increase or decrease flow. The amount of eccentricity will have a knock-on effect on the tooth length, and can weaken the gear. If the eccentricity is changed then the lobe radius of the outer diameter must also be changed to reduce the stress in the lobe.

The number of teeth chosen will have a major impact on pump performance. Fewer teeth mean that the overall diameter of the pump can be smaller, which in turn makes it easier to package in the engine. Fewer teeth also result in a reduction in power loss for a given flow rate, but the downside is that they have to run at a higher rotational speed, which can increase wear.

On the other hand, more teeth will give fewer pressure fluctuations and hence a smoother overall oil pressure. The fluctuation is sometimes referred to as pressure or flow ripple, and is an unwanted side-effect of any positive-displacement pump. The effect of pressure ripple is mechanical vibration in the oil circuit in the engine, which can damage components such as oil hoses, seals, and bearings, so any measures that can minimise pressure ripple are likely to be of great benefit.

External Gear Pumps

External gear pumps consist of two gears laid side by side, with one gear driving the other. The movement of the oil by an external gear pump is similar to that for an internal gear pump, and like the internal gear pump it is the teeth on one gear that drives the teeth on the other.

When the gear teeth come out of mesh they create an expanding volume that is filled by the oil fed from an inlet port. This oil is trapped in the gap between the teeth and the inner diameter of the housing, and as the gear rotates then so too does the oil. The gear teeth then come back into mesh, collapsing the volume of oil and forcing it into the outlet port and out of the pump. Most race engine oil pumps contain spur gears with parallel teeth, but in other applications the gears can be angled in a chevron pattern.

The tooth profile for an external gear pump is different from that used for an internal gear pump. Most external gear pumps contain gears with involute tooth profiles. An involute curve is the spiralling path created by unwinding an imaginary taut string from a stationary circle. The curve is prescribed by a number of inputs, including module, pressure angle, addendum modification coefficient and base circle diameter. All these variables can be adjusted to give the final tooth shape.

External gear pumps will typically have supporting journal shafts at either end of the rotors (as do internal gear and external lobe pumps for that matter), and the shafts run inside bearings or bores in the pump housings and cover. Because the shafts are normally immersed in oil, it is possible to avoid the need for bearings, and in some pump designs they can run directly in the bore so long as the size, cylindricity and positional tolerances of the bores are carefully controlled during manufacture.

However, wear in the journal bore can still sometimes be an issue. As a result, some pumps will have bronze bushes in the housing and cover plate. Further, some pump designs have small drillings to help feed oil from the high-pressure outlet port to the centre of the journal bore for lubrication purposes. That might have a small effect on pump efficiency but it can prolong the pump’s life, as wear in the journal bores can cause the rotating gears to start to lose alignment after an extended period of running.

External Lobe Pumps

A popular solution for scavenge pumps is the external lobe pump, which operates in a similar manner to the external gear pump. The external lobe pump consists of two identical rotors rotating in opposite directions, with lobes that do not contact one another.

They are timed and driven by a pair of gears that are mounted on the same shafts as the lobes, which can be a disadvantage as extra room is needed to house the timing gears. As the lobes come out of mesh they create an expanding volume in the inlet port.

The movement of oil in an external lobe pump follows the same principle as in an external gear pump. The lobes rotate to create an expanding cavity that is filled with oil. The oil rotates around in the lobes until the lobes mesh again, forcing the oil into the outlet port and out of the pump.

External lobe pumps are sometimes referred to as a Roots pump, named after the American inventors and brothers Philander and Francis Marion Roots, who established the Roots Blower company in the 19th century. The lobe concept used in a Roots-type air induction blower is similar to that used by external lobe oil pumps.

External lobe pumps make an excellent choice for scavenging oil out of the engine because they are capable of pumping much larger volumes than could be attained with an external gear pump. This extra pumping capacity means they will also remove the crankcase gases, which creates a beneficial depression in pressure within the crankcase. In addition, because the lobes don’t contact one another, they don’t suffer from wear damage when pumping just air.

Early rotor lobe profiles were generated from a series of radii, but now more complex curves can be constructed to optimise the amount of volume displaced by the meshing lobes. The profiles of the lobes tend to be a mixture of epitrochoid and hypotrochoid curves used in internal gear pumps.

The number of lobes is another source for development. Traditionally, scavenge pumps would have twin lobe rotors, but there has been a trend to increase the number of lobes to three or more per rotor. While two-lobe rotors will have lower frictional losses than multiple-lobe rotors, having more lobes can help the rotor withstand the ingestion of debris.

As already discussed, the disadvantage with any positive-displacement pump is that the teeth come in and out of mesh, resulting in a variation in output pressure. On a lobe pump, one way to alleviate that is by using helical lobes that are twisted rather than straight. Although such helical gears can promote axial forces on the rotor, they have been proven to give a smoother flow of oil at steady speed.

Lobe pumps can be very efficient as they can offer a much larger ratio of displaced volume per unit volume than their geared counterparts (in effect, the volume between the lobes is larger than the volume between the gear teeth). However, there is a chance of more leakage in a lobe pump as the route back between the rotors is less convoluted than that in an internal gear pump.

This is particularly true with a twin-lobed rotor, so the clearance from the tip of one lobe to the base circle of the other needs to be as small as possible. That can only be achieved with precision machining of the rotor profiles and extremely accurate positioning of the rotor bore and journal bores in the housing and cover.

Pump Efficiency

There are a couple of useful parameters to describe how the efficiency of an oil pump: volumetric efficiency and overall efficiency, which are both expressed as a percentage. Both have to be obtained from empirical tests that are normally carried out on a test rig. Typical values of volumetric efficiency and overall efficiency vary markedly depending on the type of pump, and will also change with speed and pressure.

Volumetric efficiency is defined as the ratio of the actual volumetric amount of fluid moved by the pump compared with theoretical volume of fluid that should be moved. The actual volumetric flow rate has to be measured with a flow meter, and it is important to include the flow through the pressure relief valve if one is present.

Complex equations and CFD (computational fluid dynamic) simulations can be derived to determine the theoretical flow rate, but there is a simpler approach that will give a good approximation. First, measure the volume of the gear tooth space; normally this can be determined from a CAD model. The theoretical volumetric flow rate is then the product of the tooth space volume, the number of teeth and the rotational speed of the gear.

The overall efficiency is the ratio of the hydraulic power created by the pump relative to the (input) power required to drive the pump. The input power has to be measured, normally by fitting a torque sensor to the pump drive, and the input power is then the product of the measured torque and the rotational speed of the pump drive. The hydraulic power is then the product of the measured flow rate and the pressure rise delivered by the pump. 

The goal for most race engine oil pump designers is to minimise any losses in the amount of hydraulic power generated by the pump. Some of these are attributed to the journal shafts running in the housing, cover and seals.

Further, in a gerotor pump, most of the mechanical losses come from viscous drag on the outer diameter of the outer gear and its end faces, hence such pumps will sometimes incorporate low-friction coatings in the bore and faces of the housing and cover. In a gerotor pump, reducing the outer diameter of the outer gear can be very beneficial to reducing viscous drag.

The clearances between the inner and outer rotors and the housing will also have a large effect on both volumetric and overall efficiency, so the lengths and diameters of the rotors and housing bore have to be tightly controlled. For example, most gerotors for oil pressure pumps will have clearances of around 0.1 mm radially between the outer rotor and the housing, 0.05 mm axially between the rotors and the housing, and tip clearances of around 0.08 mm.

These clearances are compounded by tolerance stack-up of multiple parts. For example, to achieve the axial clearance of 0.05 mm the tolerance on the length of rotors and the depth of the housing bore have to be around 0.025 mm per part. Minimal clearance will mean less loss due to leakage of the oil across the end faces from the pressure side to the suction side, but too little clearance can cause binding of the rotor. It is also possible to have debris trapped between these end faces if there is insufficient clearance.

Pressure Control

Undesirable thrust loads on the rotor end faces can arise if there are excessive fluid pressures inside the pump, resulting in sometimes catastrophic seizure of rotors. Most pump manufacturers have their own design methods to avoid this build-up of back-pressure, with small ports and galleries helping to drain the oil from the faces.

Some oil pump designs include small, shallow pockets in the end faces of the housing near to where the gears mesh. These are intended to relieve the lateral forces created between meshing gears, and which would otherwise cause the rotor and journals to start to bind and seize in their housing bores.

Also, most oil pressure pumps will incorporate a pressure relief or regulation valve to prevent over-pressurisation (you can read our technical blog on pressure relief valves here). As the output pressure is proportional to the pump speed, the valves are designed to operate at a pre-selected engine speed.

Most pressure relief valves consist of a spring-loaded piston. The spring rate and length are chosen so that the piston moves when the oil pressure reaches a desired value, and the piston will then slide to open a gallery to divert oil away from the high-pressure outlet, thereby capping the rise in pressure.

When designing pressure relief valves, consideration has to be given to where the by-passed oil will return to. Some designs will return the oil back to the pump inlet, whereas others will recirculate the oil back into the engine. Feeding oil back into the inlet can increase the oil temperature inside the pump and can also aerate the oil, which are both undesirable side effects.

Material Selection

As with any critical engine component in a race engine, the choice of material for the individual components in an oil pump needs careful consideration. While they may not be subject to the loads and stresses seen in other engine parts, oil pump housings still have to be able to withstand the stresses and deflections induced by the rotating internal components. And although a cracked oil pump housing or cover might not be catastrophic to start with, the consequences are still serious, as a gradual leakage of oil or reduction in oil pressure can eventually lead to engine failure.

Oil pump housings and covers are therefore normally made from a very high-grade aluminium alloy, with careful design to ensure that the rotor and journal bores remain as rigid as possible under all operating conditions. Aluminium is also an excellent material to machine, which can make it easier to achieve the very tight machining tolerances required. Some pump housings and covers are also anodised for protection against the harsh environments they find themselves in.

The rotors for both an internal and an external gear pump will be subject to very high contract stresses when the gear teeth mesh. Consequently, quite often they are made from high-grade steels, similar to gears from other drive systems in the engine.

There is however a fundamental problem when running steel rotors in an aluminium housing, and that is the different thermal expansion rates of the two materials. Aluminium has a thermal expansion coefficient that is more than twice that of steel, meaning that as the oil pump temperature increases, the housing will grow more than the rotors, which will lead to increased clearances, increased leakage and consequently a reduction in efficiency.

Some pumps will feature bronze or brass as the rotor material, as these materials are closer to aluminium in thermal expansion rates. Conversely, some pumps with steel rotors will have cast iron housings that have similar thermal expansion coefficients.

The rotors in an external lobe pump can enjoy a slightly easier life then their geared counterparts because the lobes don’t come into contact with one another, so there is no contact loading to worry about. This means that wear isn’t an issue on the lobes, and the materials used can be as light as possible, for example aluminium alloy or even carbon fibre. However, caution has to be applied to the wear properties of the lobes as they will have to cope with any debris sucked in from the crankcase, hence the need for good filtration on the scavenge pump inlet.

Coatings are another option that manufacturers can turn to, either to combat wear between the meshing gear teeth or to reduce friction between running surfaces, particularly between the end faces of the rotors and the flat faces of the housing and cover. PVD-applied coatings such as DLC can help to reduce wear, and PTFE-type coatings have been used over the years to reduce friction.

This feature on crank bearing lubrication is based on an article written by Modatek’s Matt Grant for Race Engine Technology, issue 105. You can purchase a back copy of this publication here: https://www.highpowermedia.com/Product/race-engine-technology-issue-105

One of the critical features of a con rod, and one that is often overlooked, is the joint face. This is the pair of mating faces on the rod and cap halves, and it plays an important role in keeping the con rod together.

In an ideal world there would be no rod and cap halves, and indeed some engine configurations consist of a crankshaft assembly that can be built up from multiple separate sections of crank pin journals and webs. Such an arrangement would then allow for the con rod and big-end bearing to be slid onto the crank pin journal. However, the practicalities of packaging and assembly mean such layouts are only suitable for engines with fewer than four cylinders.

Most race engines will comprise a single-piece crankshaft containing crank pin journals separated by webs and counterbalance weights. So that it can be assembled onto the journals, the rod for each cylinder must consist of a rod half and a cap half.

In most applications, the rod and cap halves both contain plain bearing shells (or in some cases needle roller bearing cages), and these two halves are then bolted together to effectively clamp the bearing. The mating faces of the rod and cap halves are generally termed as the joint face or split line.

Joint Face Design

The correct design of the joint faces of the rod and cap halves is fundamental to the performance and longevity of the con rod assembly. Failure of this joint face will almost certainly lead to a failure of the big-end bolts or bearings, which normally results in a catastrophic engine failure.

Con rod designers will therefore pay close attention to the shape of the joint face, making sure it can withstand both the pre-loading from the big-end bolts and the cyclic loading during engine operation. The design of the rod and cap halves around the split line will have a fundamental effect on the stresses seen by the big-end bolt during these cyclic loads that come about as a result of the reciprocating forces pulling and pushing on the small end of the rod.

A good rule of thumb is that if the stiffness of the housing material around the big-end bolt and at the joint face is increased, the proportion of the cyclic load that the big-end bolt will experience will be reduced. Simply increasing the stiffness around the joint face by adding extra material however isn’t always possible owing to the limited space available – and besides, it will result in a detrimental effect on big-end mass, where every gram matters.

Remember that for static balancing of most crankshaft configurations, every gram on the bottom end of the rod multiplied by the radius of rotation will need a corresponding amount of mass multiplied by radius on the crankshaft balance weight. Any weight saving that can be applied to the bottom end of the rod therefore has a twofold reduction in rotating masses.

Increasing housing stiffness has the beneficial consequence that the bolt size can be reduced, which in turn means the overall size of the joint face can come down. This can lead to an iterative design process to optimise the bolt size and housing stiffness, which is normally carried out either by time-consuming and expensive trial and error or 3D finite element analysis (FEA).

Another key parameter that has to be optimised when designing the joint face is to try to bring the rod bolts inwards and as close to the big-end bore as possible. The major restriction here is the wall section at the joint face between the holes in the rod and cap halves and the big-end bore.

Keeping the big-end bolts as close to the bore as possible makes for a lighter and more compact con rod bottom end, and can reduce the amount of bending load seen in the bolts. Unlike some other classical bolted joint scenarios, big-end bolt loads are not just axial but are a combination of axial and bending loads thanks in part to the oscillatory motion of the bottom end of the rod.  

Knowing that pressure is force divided by area, it is clear that the surface pressure seen at the joint face of a rod is greatly influenced by the area of the face. Increasing the joint face area normally means that the surface pressure is lowered, and relatively simple calculations can be derived to determine this pressure as a function of gas and inertia loads combined with the pre-load from tightening the big end bolt.

The effect of housing stiffness and deformation can complicate matters somewhat, which is where FEA can be used to predict surface pressures under different engine operating conditions and at different stages in the two- or four-stroke cycle. There is also the variation in surface pressure across the joint face to consider: the surface pressure can be higher closer to the bolt holes, and is reduced near the big-end bore as the stand-off of the big-end bearings is effectively trying to push the joint faces apart (plain bearings are deliberately designed to be slightly bigger than the housing bore in the con rod so that they are clamped in place when assembled to prevent movement).

Some con rod designs try to minimise this surface pressure variation by adding a very shallow groove or channel in the joint face, running in line with the bolt hole and parallel to the big-end bore axis. This relief can help to raise the pressure at the joint face near the big-end bore to combat the bearing stand-off forces.

Manufacturing the joint faces on both rod and cap halves can be a relatively simple operation if the faces are intended to be flat, but the geometric tolerance of the joint faces with respect to the big-end bore is important, as is the final surface finish.

The joint faces in both halves must be machined before the big-end bore is finished, which can complicate the specification of geometric tolerances; as such, specifications need to pay due regard to the sequential stages of the manufacturing process. Traditionally, joint faces are ground or lapped to achieve a surface finish of around 0.4 Ra. A mirror-like finish isn’t necessary, but conversely a roughened surface is also to be avoided.

Rod & Cap Location

In most cases the rod and cap halves are machined from a forging or billet blank that at some point has to be cut along the split line to allow the joint faces to be machined. Sometimes the rod and cap are actually machined separately, which would be the case if they were made from different materials. Once the joint faces have been machined, the rod and cap halves are then clamped back together with con rod bolts (sometimes referred to as slave bolts) so that the big-end bore can be machined.

It is imperative that after the big-end bore is finished, every time the rod and cap halves are subsequently separated and reassembled they are aligned exactly as they were when the big-end bore was machined. Also, some rods are crank-guided, meaning that they have thrust faces on both sides around the big-end bore to control the lateral movement of the rod along the crank pin journal, and the faces need to remain in their as-machined condition after reassembly.

Consequently, the need for the geometry of the big-end bore and thrust faces to remain as close as possible to how they were machined means there must be some method of alignment between the two halves. Many steel and titanium race engine con rods use ring dowels for this alignment.

The ring dowel, or alignment sleeve as it is sometimes called, is pressed into a counter-bore that has been machined into the rod bolt hole in one half, and the dowel then aligns with a matching counter-bore in the other half. Usually the dowel itself is made from a steel alloy, hence both it and (in the case of steel and titanium rods) the surrounding material don’t distort too much when there is a small amount of interference.

This means that repeated removal and assembly of the two halves during manufacture and engine build still gives a repeatable amount of realignment. With an aluminium rod and/or cap half, a steel ring dowel is likely to distort the softer surrounding material, so a serrated joint face is favoured.

One disadvantage of the ring dowel design is that it means having a bigger distance (pitch) between the centres of the big-end bolts, otherwise there won’t be enough wall thickness between the counter-bore on each side for the dowels and the big-end bore. Deleting the counter-bores for the dowels allows the pitch of the bolts to be brought in, which will reduce the bending loads on the bolts, as discussed.

It also allows the width and length of the joint face to be reduced, which in some applications is favourable for packaging the big-end of the rod in confined spaces. That is particularly so in vee engine design, where the designer wants to minimise the bank stagger (the lateral offset distance between the bores on either bank), which brings a knock-on reduction in overall engine length.

Alternatively, a dowel pin on both sides and next to the bolt holes can be used for alignment, and indeed has been a proven robust method for the alignment of rod and cap halves for decades. In an area of the engine where mass is at a premium, some con rods have dowel pins that are less than 4 mm in diameter. Not only does this result in a lighter pin, there is also less material required around the holes in the rod and cap halves. The centre point of the dowel pin hole is carefully chosen to minimise the width and length of the joint face for the reasons already mentioned.

The holes for both ring dowels and dowel pins are relatively easy to machine, and can provide a very good level of realignment tolerance. Traditionally, such holes are slot-drilled first, followed by a reamed hole that can give an ISO H7 tolerance band (in the case of a 4 mm hole, an H7 tolerance band is 12 microns, enough to give a suitable amount of interference with the dowel pin.)

Commercial ring dowels and dowel pins are readily available from numerous suppliers. Dowel pins are normally made from hardened steel or even silver steel in some cases, and the outer diameter is precision ground for excellent roundness tolerances along the entire length of the pin. Sometimes in the case of a solid dowel pin there is the option to include a small, flat section along its length to allow air to escape when the pin in pressed into the blind hole, although this is less common with smaller dowel pins.

Coiled spring pins (sometimes referred to by the tradename Spirol) are an established alternative to the solid variant. They are produced with sufficient control of their diametric tolerance to enable them to align the rod and cap halves but their construction makes them able to absorb shock loads.

Another possibility is to include a shoulder on the big-end bolt, with the shoulder diameter carefully controlled to give a slight interference in the holes in the rod and cap halves. This is only really practical through when the bolt is used in conjunction with a nut.

Joint Face Serrations

On a conventional flat joint face, in the absence of sufficient pressure, the rod and cap halves can slide, albeit only very slightly but enough to result in fretting damage. But there has been a growing trend towards a new type of joint face: serrated faces, like those on our Cosworth CA2010 con rods.

If machined accurately, serrated faces will be in contact on the flanks of the serrations, restricting any lateral movement and providing a rigid and secure joint. This is particularly important for aluminium con rods, as aluminium is prone to fretting. The flanks of the serrated faces will also have a larger surface area than that for a flat face, which results in a greater contact surface area.

These serrations can increase the housing stiffness around the big-end bore, which can have a beneficial effect on reducing the ovalisation of the big-bore during engine running and can also reduce bending loads on the big-end bolts. The big-end bore will distort into an oval shape when the piston approaches TDC, which can cause wear between the back of the big-end bearings and the con rod big-end bore surface, and create issues with the oil film breaking down between the crank pin journal and the inner diameter of plain bearings.

Initially, only straight-cut serrations were available, with the serrations all running in a direction parallel to the big-end bore axis. While that would give good location in one direction, dowel pins or ring dowels would still be needed to give complete alignment, to ensure that the thrust faces on the rod and cap halves were lined up accurately.

Some modern race engine con rods include curved serrations on the joint face that can align the rod and cap halves in all directions, making dowel pins or ring dowels obsolete. The path of the curvature of each serration can vary, but most con rod manufacturers use concentric serrations that are centred at a theoretical point outside the rod. Each side has to have a different centre point, otherwise the rod and cap halves would still be able to rotate about this point, so the designer has to make sure that all degrees of freedom are fixed.

There are numerous ways to machine the serrations on the joint face, including grinding, wire cutting, EDM (electrical discharge machining, also known as spark eroding) and even conventional milling. The actual profiles of the serrations vary between the different con rod manufacturers: some favour more than ten small serrations on each side, others opt for just two or three larger serrations.

The flank angles are nearly always at 45º to the split line to ensure an even distribution of stress in both the rod and cap halves. Understandably, the positional accuracy of the serrations is extremely important if the rod and cap halves are to mate together correctly.

Fracture Splitting

Although more common in the high-volume world of automotive con rods, fracture-split rods are now starting to appear in some race engines. Fracture splitting has been an established process for several decades, and while the method for splitting the rod into the rod and cap halves might seem crude, it actually requires a lot of careful design and process control.

Essentially, a defect is intentionally added to what will become the split line and, with the big end supported, the rod is struck hard enough to split it into two halves. This creates a very high strain rate on the split line, which is enough to make the normally ductile material behave in a brittle manner. This brittleness is necessary as there can’t be any deformation at the split line, so the resultant roughened surface of the joint face is made up of a vast number of peaks and troughs that will locate perfectly when the two halves are reassembled.

This repeatable accuracy in realignment is what makes fracture splitting so appealing for mass production. It means there is no need for dowel pins, ring dowels or serrated faces, thereby reducing cost. It also removes the need to machine or grind the joint faces, which provides another useful cost saving.

The drawback with fracture splitting, and perhaps the reason why it hasn’t become more popular with race engine con rods, is that it is only suitable for certain materials. The rod needs to be made from a material that has a sufficient strain rate sensitivity to allow it to fracture in a brittle manner. Road engine rods that are fracture split are often made from high-carbon wrought steels, as the high carbon composition will lower the content of ductile phases in the steel, but such materials aren’t always suitable for race engine rods.

Recently, some con rod manufacturers have been able to introduce fracture splitting to titanium rods. Titanium might not seem an obvious candidate for fracture splitting because it has a high amount of ductility, especially when compared to most steels. The elongation (a measure of ductility) of titanium alloys that are prevalent in race engine rods can be around 20%, versus around 13% for steel alloys, which means that in most cases titanium is more resistant to fracturing than steel.

The trick to fracture splitting is to find a way to reduce the fracture toughness in the split line zone only, without a detrimental effect on the rest of the rod, where high strength and good elongation properties are still important.

Irrespective of the material chosen, there will also be some sort of feature added from which the crack will initiate, most commonly a small vee-shaped groove or notch where the split line is required. Such notches will allow the crack to propagate and, in the case of the example from Yamaha, move along the grain flow lines that were originally created during the extrusion process.

The final method to weaken the area around the split line is to freeze the con rod in liquid nitrogen, which can take the temperature down to below -200 C. All metals go through a brittle-to-ductile transition as the temperature is lowered (technically this should be termed ductile-to-brittle), and the temperature at which this transition occurs depends on the composition of the metal.

Angled Joint Face

Alongside the radius of the crankshaft web or counterweight, the locus (path) of the heads of the big-end bolt as they move through one complete engine revolution can dictate the height of the sump. A shallow sump can have many benefits for the packaging of the engine, such as enabling the engine to be positioned lower in the car to lower the position of the centre of gravity for better vehicle performance.

In most cases the split line of a conventional con rod will be perpendicular to the rod’s axis, and cut-outs can be added to the engine architecture to clear the rod bolt locus. However, in the quest for better packaging of the bottom end of the engine, some engine designers look to angled joint faces. Angling the split line on the rod can also help to raise the rod bolt locus, thereby permitting the bottom of the sump to move up.

Choosing an angled joint face for a con rod can impose restrictions on the design of the threads in the rod and in the assembly of the big-end bolts. If the rod requires an angled joint face then it is almost certain that the thread will need to be in the rod half, as the big-end bolt will only be able to pass through from the cap half. If the angle of the joint face is severe enough then one of these tapped holes in the rod half will have to be a blind hole.

Although that isn’t necessarily a major problem, it does mean it won’t be possible to measure the stretch of the rod with conventional tooling; in these circumstances, the rod bolt will have to be tightened to an angle instead. As discussed, only a small proportion of the applied torque goes into the pre-load of the bolt, so the rod bolts aren’t normally tightened to a torque figure. To get an accurate and repeatable amount of pre-load, the bolts are normally tightened up to a calculated amount of bolt stretch.

This feature on con rod joint faces is based on an article written by Modatek’s Matt Grant for Race Engine Technology, issue 109. You can purchase a back copy of this publication here: https://www.highpowermedia.com/Product/race-engine-technology-issue-109

Turn your volume up to 11 and sit back and enjoy 90 seconds of pure, unadulterated noise in this Cosworth Formula 1 engine test as a CA2010 engine is pushed through its paces on Cosworth’s transient dyno at speeds of up to 18,500 rpm.

The Cosworth CA engine represented a pinnacle in the design of high-speed naturally aspirated engines for Formula 1 racing. Calling on over 50 years of experience that dated back to the legendary DFV, Cosworth created their most ground-breaking engine ever, capable of reaching 20,000 rpm in 2006.

Sadly, in 2010 the engine speeds were restricted by regulation to 18,500 rpm, but that didn’t stop the cars of that era sounding amazing.

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.

If you enjoy listening to the roar of a Formula 1 engine being put through its paces then you can download it as a ringtone for your phone (Android only, sorry!).

The Need for Speed

In the ear-splitting ‘90s and 2000’s, engine speed was THE metric that Formula 1 engine manufacturers wanted to push to the limit. In a naturally aspirated engine, the goal is to get as much fuel and air into the combustion chamber, and the best way to do this is to run the engine as fast as possible. Unlike power figures, engine speed is something that rival competitors can measure from the side of the track, using audio recording equipment.

It is generally acknowledged that BMW were the first F1 engine manufacturer to break the 19,000 rpm limit in 2002. Toyota soon leapfrogged their fellow German rivals, reaching 19,200 rpm in 2005 despite the introduction of rules to increase engine mileage. These regulations caused a brief respite in the push for speed, but by the end of the V10 era in 2005 the entire grid was running at 19,000 rpm.

Switch to V8 Power

In an attempt to curb engine power, the FIA mandated a switch from 3.0 litre V10s to 2.4 litre V8s in 2006. However, this had little or no effect on the push for speed. Cosworth’s Bruce Wood explained how they did this to Race Engine Technology as part of their exposure on the CA . “To go faster you just have to keep making the bore bigger and the stroke shorter and sort out your valves.

“While developing the TJ we did tests on our single cylinder rig of 96, 97 and 98 mm bores – it was all about higher speed. We were considering a bigger bore and bigger valves and a compound valve angle before the mandatory switch to V8s was brought in. We had thereby established that combustion was OK with the 98 mm bore (the maximum permitted in 2006), so there was no reason not to move to it…”

In 2010 Cosworth returned to Formula 1 with the CA2010. By this time the rules had necessitated a cap in speed of 18,500 rpm. Nevertheless, the cars still sounded incredible, and drop in speed didn’t prevent engines like the CA2010 from reaching power figures of over 775 bhp.

Keeping It Together

20,000 rpm is a huge achievement – as this Formula 1 engine test video shows, even 18,500 rpm sounds terrific, but running at these speeds results in a number of engineering headaches. First of all, the inertia loads on the moving components roughly increase with the square of the speed. So a jump from 10,000 rpm to 20,000 rpm might be double the speed but will result in a quadruple increase in inertia loads.

The second problem with increasing engine speed is that there is a corresponding increase in vibration. This was compounded with the switch from V10 to V8 engines in 2006. A V8 engine like the CA equipped with a flat plane crankshaft will naturally have out-of-balance vibration in a horizontal direction.

Bruce Wood gives a further insight into the problems that this horizontal shaking of the engine would cause on the CA. “When we first started running the CA, the scavenge pumps, which are held onto the sump with horizontal bolts, would fall off. Those are 8 mm cap screws, the heads of which snapped off because of the unbalanced force, which is why our scavenge pumps are now secured by Multiphase bolts!”

Torsional Control

One other problem to resolve before unlocking 20,000 rpm is torsional resonance of rotating components like the camshafts and crankshaft. Torsional resonance has always been an issue in Formula 1 engines – Keith Duckworth had to resort to flexible compound gears to keep the gear drive intact in the DFV in 1967. Fast forward 4 decades, and double the engine speed, and the torsional problems become a whole lot worse.

“In terms of the torsional vibration inside the engine, we knew what we were up against, which is why the CA has far more damping devices in it than our previous V10 engines.”, says Wood. “We have a ‘compliant’ gear train that has been in our Formula One engines for years, then in addition (to two dampened compound gears) the CA has compliant quill drives within each of its two auxiliary drives, a big viscous damper on the back of the crankshaft, viscous dampers on the back of each camshaft and friction quill dampers in the front of each camshaft. That means in total it has 13 dampers – 14 when fitted with KERS.”

Modatek provide parts and consultancy for a wide range of Cosworth historic engines, including the CA. Want more information? Just send us a message through our Contact Us page.

To the untrained eye, the skirt of a piston would appear to be perfectly cylindrical. To be fair, even to the trained eye that would seem to be the case. But if you were to roll a piston on its side down a flat slope, you’d notice that it gradually wanders off path – that’s because the shape of the skirt isn’t round! In this technical blog we uncover what shape it actually is, and why.

Looking at the piston side on, the profile of the skirt is actually barreled, not straight. And looking from the top, the shape of the piston is oval, not circular. Actually, you’d be hard pressed to notice this by eye, as we are talking in terms of fractions of a millimeter, but its enough to make a difference when the piston is running in an engine.

This barreling and ovality is intentional – it’s there to take account of the deformation of the piston when it is loaded up and heated up, which has the potential to result in inconsistent bore clearance and possible seizure.

Bore Clearance

The clearance between the skirt and the cylinder bore’s inner diameter is critical, for many reasons. Too much clearance will allow the piston to tilt in the manner described above, whereas insufficient clearance can result in catastrophic seizure. Note also that tilting the piston can lead to increased movement of the rings and more wear of the ring grooves.

Generally, the running clearance is dictated for full-load operating conditions. The clearance is influenced by engine speed and piston temperature, so other operating conditions may also need to be considered. For example, the piston temperature and hence expansion can increase during lean running, which can lead to an unforeseen reduction in clearance.

Note that the recommended bore clearance is also affected by the materials of the piston, cylinder liner (if present) and cylinder block, specifically with regards to dissimilar thermal expansion rates. Even the two grades of the most prolific aluminium alloys in use these days – 2618 and 4032 – have different thermal expansion rates that mean the bore clearance needs to be different for each alloy.

Normally a piston made from a 2618 alloy will require more bore clearance than one made from a 4032 alloy. That isn’t an issue during running, as once the pistons have reached their operating temperatures the chosen bore clearances will be the same for both alloys, but it can cause a problem on start-up or when cold as the extra clearance required for the 2618 piston can create more noise.

Piston Skirt Loads

In a conventional reciprocating four-stroke engine the small end of the con rod will articulate forwards and backwards around the gudgeon pin as the crankshaft completes one revolution. As a result, the piston will be subjected to side loads that vary for each of the four piston strokes.

The largest side load by far will be during the power stroke, when a combination of inertia and gas loads will push the piston sideways as well as downwards. This is commonly referred to as the major thrust side load, and its direction is always on the side that is against the direction of rotation.

For example, if the crankshaft rotates clockwise when viewed from the front then the major thrust side will be on the left-hand side. The opposite side of the piston is referred to as the minor thrust side, and the minor thrust side is subjected to a lower sideways force during some of the other stages of the four-stroke cycle.

As well as deforming the shape of the piston, these major and minor 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 can exasperate 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 even be detected and heard outside the engine, commonly called piston slap.

Piston Skirt Deformation

The external loads imparted to the piston from the gas pressure in the combustion chamber above the piston are immense, and will cause strain deformation of the piston shape and in particular the piston skirt area. The resultant shape of the skirt after this deformation is a complex one.

During the firing cycle the crown will be deformed inwards by the loads from the gas pressure in the combustion chamber, effectively splaying the skirt outwards. If the piston were shaped like a cup then the skirt deformation would be circular, but of course the piston incorporates the pin boss and structural ribs, so the deformation isn’t even.

The pin bosses provide support along the pin axis, so the skirt deformation at the piston’s open end can be broadly described as oval, with the diameter increasing along the pin axis and decreasing at the major and minor thrust directions. The shape isn’t completely elliptical though, as there are additional strain deformations in the skirt to consider. The pin bores themselves are subject to loads from the gudgeon pin, which can effectively flatten the skirt, and the skirt itself is deformed by the contact pressure acting on the skirt from the cylinder bore.

Superimposed on top of these strain deformations is the effect of thermal expansion. The temperature is highest at the crown, so at working temperature the diameter around the crown and around the lands will expand outwards more than the diameter of the skirt.

There will also be a gradual reduction in diametric expansion down the length of the skirt, decreasing as one goes down the skirt away from the crown. Like the strain deformation, the thermal expansion of the skirt isn’t even due to the presence of the pin bosses and structural ribs.

When the strain deformation and thermal expansion of the piston are combined, the resulting skirt topography ends up being a combination of broadly oval sections that increase and decrease in size down the length of the skirt, following what is best termed a barrel profile. Note that some manufactures refer to the barreling as cam or curvature.

Now, if one assumes that the bore in the cylinder block or cylinder liner is perfectly cylindrical during running (which it isn’t) then the shape of the piston skirt’s profile when it is machined has to be carefully designed so that it deforms (both by strain deformation and thermal expansion in the manner described above) when running to create a complementary perfect cylinder. As a result, piston designers will have to specify the necessary ovality and barreling of the skirt and lands that the machinist will need to achieve at room temperature.

Piston Skirt Design

When designing a piston, special attention is paid to the topography of the skirt. The shape of the skirt can have a large influence on engine performance, as it can control the tilting of the piston, lateral displacement, oil film thickness and frictional losses. As mentioned, the shape of the skirt is defined by barreling and ovality.

The simplest design of piston skirt will have the same amount of ovality along the entire length of the piston, and in some cases this will suffice. The designer ensures that the area of the skirt that will have the largest amount of expansion is catered for, but the remaining sections of the skirt will have a looser fit during running than might be desired, which can cause problems with reduced stability of the rings and poor guidance of the piston in the bore.

Going one step further to solve these issues, the ovality at staged sections down the length of the piston can be prescribed as a series of ellipses, which in turn are defined by their major and minor diameters, and which change in size following a barrel form. Commonly, this is done at discrete points roughly every 1-2 mm down the length of the piston.

The derivation of these ovalities would traditionally have been found from trial and error, requiring a series of engine tests to monitor the condition of the skirt. The pistons could only be inspected by removing them, hence such testing would require a number of rebuilds.

Too much contact between the skirt and the wall of the bore would be indicated by scuffing, which would indicate the areas that needed to be addressed. In the worst case, overly excessive contact would cause a catastrophic seizure, making it impossible to inspect the running surfaces left in the piston debris.

With the advent of FEA (finite element analysis), piston designers can calculate the combination of strain deformations and thermal expansion to derive the required skirt profile. Load cases for various stages in the operating cycle and different running conditions can now be considered to optimise a single profile for a number of scenarios. In addition, the piston designer can now take into account the effect of bore distortion.

The assumption that the bore is perfectly cylindrical during engine running isn’t necessarily true, even if the bore has been honed at temperature and with torque plates fitted to simulate distortion from the assembly and running loads induced by the cylinder heads, sump or mains caps, and crankshaft mains journals. There will still be some distortion to this circular section, which normally corresponds to the locations of the studs or fasteners next to the bore.

FEA can be a very powerful tool to predict the actual shape of the cylinder bore, and the results from it can be used to define the required piston skirt profile. Piston designers are turning to increasingly intricate computer-derived simulations, creating complex dynamic multi-body simulations based on the structural and thermal FEA of the piston and bore. The results of this analysis can create skirts that have the optimum contact conditions for a wide range of operating conditions.

So, the first objective of a good skirt design is to ensure that there is the correct clearance between all the parts of the lands and load bearing area of the skirt and the wall of the cylinder bore. As we’ve already seen, there are a number of factors that can influence bore clearance, and the chosen profile must take into account each of these factors.

Nowadays though, designers are also considering the dynamic behaviour of the oil that is trapped between the skirt and the bore. There has been plenty of published research into the effects of piston skirt profile on the lubrication of the piston and its subsequent influence on friction.

Such research concludes that the frictional behaviour of the piston is heavily influenced by the mode of lubrication (hydrodynamic, mixed and boundary), which in turn is constantly changing throughout the four-stroke cycle owing to the reciprocating motion of the piston. Further, it has been noted that almost 80% of the frictional power loss from the skirt occurs during the power stroke.

Piston Skirt Machining

The required shape of the skirt by the designer can be machined by mounting the piston on a rotating spindle, which is then moved in and out of a rotating grinding wheel whose movement is controlled by an eccentric cam. Manufacturers use specialist CNC piston turning centres for this, and they can be programmed by entering the required diameters and ovalities at the discrete sections down the length of the piston, as defined by the designer. Note that this type of machining will extend up above the skirt to the lands, and can include the ring grooves themselves.

More advanced machining centres now use modern diamond turning instead of grinding to achieve a tighter tolerance on the skirt profile. Some machines will need to be temperature controlled or will have an element of temperature compensation built into them to eliminate the potential problem of thermal expansion during machining – even a shift of 10 ^oC above room temperature can push the profile out of tolerance. In addition, as design techniques have advanced over the years the required profiles have become more complex.

As a result, skirt machining has had to be developed to allow for more complicated designs, such as the need for asymmetric skirts and even concave reverse-profile barrelling in certain sections of the piston. (On a piston with asymmetric skirts the bearing area of the major skirt is designed to be able to withstand the higher loads from the major thrust side, and as the deformation of the skirt is partly linked to side loading, it also means the skirt profile can end up being different on the major and minor skirts.)

Skirt Coatings

In theory, if the skirt profile has been designed correctly, and there is an adequate supply of oil, then there shouldn’t be any contact between the skirt and the bore, as there will always be a film of oil separating them. In reality though, that isn’t always the case – cold starts, oil contamination, overheating, extended high-speed running and many other problematic running conditions can lead to a breakdown of this hydrodynamic oil film.

Several piston manufacturers therefore apply a coating of some sort on their piston skirts, to allow the piston to survive the occasional skirt-to-bore contact. Although the major source of friction will come from the rings, there is some frictional saving that can be made with skirt coatings, and there is plenty of evidence to show that the application of coatings has helped with the durability of both the piston skirt and the cylinder bore surfaces.

Some manufacturers apply a dry-film lubricant that can be sprayed on after the skirt has been machined. For example, some of our Cosworth pistons have a fluoropolymer coating that has the tradename Xylan.

The thickness of the coating on the skirt can be quite thin, in the region of between 8 and 13 microns, and with such thin coatings there doesn’t need to be a manufacturing allowance when machining the skirt, as this coating is intended to wear (as is the case on the Cosworth TJ Formula 1 piston pictured above). Sacrificial coatings like these can also be extremely useful in identifying areas of high wear on the skirt – as long as the engine is stopped before metal-to-metal contact occurs!

This feature on pistons is based on an article written by Modatek’s Matt Grant for Race Engine Technology, issue 111. You can purchase a back copy of this publication here: https://www.highpowermedia.com/Product/race-engine-technology-issue-111

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.

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 = P_req
• 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 P_back, 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:

K_W 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