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Technical Advices > Reciprocating Components and Engine Balancing

Reciprocating Components and Engine Balancing

3.1 Connecting Rods

There is a fair amount of mystique associated with connecting rods, but one thing is certain, they are one of the most important components of an internal combustion engine. There are many opinions about what makes a good connecting rod, and I for one am not sufficiently versed in metallurgy to go about designing one. What I can do however is look to those companies that have a good track record and try and see what they do well.

Companies like Saenz, Pankl, Carillo, Crower and many others have been making connecting rods for many years, with great success. So what is so different about a specialty rod, as compared to a standard production rod. Three words describe the difference.

Fasteners - Fit - Finish.

Fasteners - In my view the most important aspect of any connecting rod is the bolts. For connecting rods that are to be exposed to the stresses of competition, only the absolute best will do. Whether you purchase the bolts you use with the rods, or buy them from after-market suppliers like SPS or ARP, the money spend on better connecting rod bolts will pay off every time.

Standard Fiat bolts are fine for normal road applications. They rate at about 90,000 to 100,000 PSI. For an engine that is going to see 8000 RPM or more this just will not do. After-market bolts start at about 160,000 PSI and go up from there. The ultimate strength of the bolt will be determined by both the metallurgy and the size of the bolt.



Stretchgauge

Due to size restrictions, the standard bolt for Fiat rods is a fine pitch threaded 8mm unit. If you are using standard connecting rods, use the 850/903/1050 units. (Remember - The 903 rods are 2mm longer) These can be cleaned up quite and then install a set of replacement rod bolts rated at 160,000 PSI. If all the components are lightened and balanced, this will give a very robust installation.

If you want even better then use connecting rods from Scuderia Topolino, Carillo, Crower or Pankl. All offer uprated bolts for their connecting rods ranging from 180,000 to 285,000 PSI. In all cases you will note that the shank of the bolt is cut down, so that it is slightly less than the original diameter of the bolt. There is a reason for this. All bolts must be able to stretch - but only a predetermined amount. Without stretch the bolt will not properly clamp the cap to the body of the rod. If the bolt stretches too much, or fails to return to its original length when loosened, DISCARD IT. It is also important "where" the bolt stretches. The reason for the shank reduction is this is the area where the stretch is supposed to take place. The diameter of the bolt in this area will be slightly less than the root, or bottom, of the threads. If this shank reduction were not there, then the bolt would have a tendency to stretch in the thread area. This would not be a good idea. In general an 8mm bolt should stretch between .005-.006 thousands of an inch at or below indicated torque. Each manufacturer will have his own specifications, and it is important that their recommendation be followed.

One thing that all of the premier companies agree on is how connecting rod bolts should be torqued and to do it correctly this takes a "rod stretch gauge". When I build a racing motor, I record the free length of each bolt and note it in the build records of the motor. If not, how would you ever be able to check whether the bolt failed to return to its free length when slackened off. The bolt is then fitted, finger tight, and torqued to 80% the specified amount. Now you check for stretch. If the bolt has stretched less than .005 increase the torque on the bolt until the required stretch is achieved, but do not exceed the maximum torque recommended by the rod manufacturer.

If you are fitting after-market rod bolts to standard connecting rods, or to rods for which they were not specifically manufactured, pay close attention the fillet radius of the underside of the bolt head. In many cases with standard connecting rods you will have to radius the edge of the hole in the rod. If you do not do this then the bolt will make contact with a sharp edge in the radius area and it is almost certain that an early and catastrophic failure will occur.

A WORD OF CAUTION - You must fully lubricate the connecting rod bolt before torqueing. This includes the threads, shank and the underside of the head. If the bolts come with special assembly lubricant, USE IT. You would be amazed at the different friction coefficients of regular motor oil, synthetic motor oil and special assembly lubricants. Also the torque recommendations will be different for all three.

One reason why I personally did do not use Carillo rods, is that the H-Beam design mandated the use of a 1/4 (approx. 6mm) bolt, where as in the Scuderia Topolino design a 5/16th (approx. 8mm) bolt is utilized. Carillo defends their use of the 1/4 inch bolt on the basis that the rod/piston combination's dynamic forces are will within the design limits of a 1/4 inch bolt. Call me a skeptic, but in the connecting rod bolt department "bigger is better" in my view.

Fit – Each aftermarket connecting rod manufacturer has their own technical solution. Carillo's H-beam design has been the standard for racing engines for many years (Even the Chinese are now making a knockoff of this design). The Scuderia Topolino connecting rod also includes an H-beam component, but this is combined with a certain aspects of the I-Beam design as well. The result is a connecting rod that has the best features of both designs and which is lighter overall..

Using a custom connecting rod allows the engine builder to vary the length as well. This freedom, along with innovative piston designs, may make impressive horsepower gains, but more important are the gains to be made in engine acceleration due to lowered rotating mass and better rod angularity. One extra consideration is that all Scuderia Topolino rods have doweled caps. This rod caps will have two hollow dowel pins to locate the cap.

Many times I am asked why I go the trouble of buying custom connecting rods. Many competitors have done quite well with the standard 850 or A112 rods, suitably reworked. I guess it depends on how much time you have and what the ultimate objective is. If you consider that you may spend as much preparing a set of standard rods, presuming you account for your own time, then buying a set may not seem so outlandish, even to the amateur competitor. By the time you do the following:

  • Grind off the excess material (large blocks of metal at the big and small ends of the rod),
  • Checked for axial twist and straightened if required,
  • Sized or honed for bearing fit,
  • Sized or bushed and honed for piston pin fit,
  • Polish the rod and then have it shot peened (stress relieving after grinding); and
  • Finally have the set balanced

then you may have spent more on your time, materials and outside machine shop labor than a custom set of connecting rods. A set of custom rods from Scuderia Topolino will have all of this done to specification.

Finish - In modern connecting rod technology there are two "finish" areas. One is the actual machine work quality, including proper alignment of oiling holes and making sure that the bolt holes have proper a proper edge radius. Finally, there is the overall finish of the connecting rod. While polishing really looks neat ( at least to the person putting the engine together), no one else is going to see all that work, unless of course the engine disintegrates and there are lots of nice, shiny polished parts that are lying around on the track. I prefer a shot blasted finish which is the finish resulting from the stress relieving process.

The second half of the "finish" equation is surface coating. Companies like Calico Coatings now have function specific coatings, including for connecting rods they have special oil-shedding coatings that assist in getting the oil off the rods and back into the pan.

If you decide that custom rods are the way to go, then the following diagram will come in handy. All of the measurements are required in order to manufacture a custom connecting rod. One measurement not shown is whether the PE is offset from the BE. I personally do not use any offset, but if you are ordering special connecting rods, you can include this specification. In addition to these measurements, if you are building a 1050 or 982cc motor, then the width of the shank of the connecting rod will also be critical. Even if the width is no greater than a standard rod, the camshaft may still have to be relieved to prevent the rod from striking the camshaft core and the sides of lobes 1,4,5 and 8.

So why would we want to lengthen the connecting rod? Well, the short answer is of course the ever present holy grail of increased horsepower. But, just what is it that we are going to accomplish with a longer rod. Here is a short list.

1. Decrease in piston skirt side force - reducing friction and reducing parasitic losses

2. Increase in Dwell Time at TDC and BDC. This allows for better combustion control, particularly at high RPM, but from a negative perspective it may affect exhaust scavenging if the head used has a poor exhaust port.

3. More equalized piston acceleration with reference to TDC and BDC - not piston speed.

For high performance engines, where the builder has already explored better porting, valve angles, exhaust systems, camshafts etc., going to a longer rod will have more plus benefits than negative ones. Certainly longer dwell time at TDC will allow for additional time for flame front travel with high RON fuels. This will allow a small increase in static ignition advance, while still maintaining the maximum pressure crank angle on the combustion stroke. Even if everything is just right, the maximum gain that one should expect from longer rods is just around 1%. In small engines this might be 1 to 1.5 horsepower.

Connecting rod angles

Many engine builders will tell you that an optimum rod angle is 1.75. Thus, the stroke/rod length combination for the 843cc motor is pretty close to ideal. Even the 903 is pretty good, but the 982cc combination is getting pretty marginal. However the 843cc engine does have a downside. By moving the piston pin lower more side load will be produced on the piston skirt. There would be a way to optimize the situation to give the 903 and 843 engines the same high piston pin location, namely using a longer connecting rod. This give the following combinations:

  • 74mm stroke - 0.995 inch (25.25mm) piston pin location - 982cc - 110mm rod length - Rod angle = 1.486
  • 69mm stroke - 0.995 inch (25.25mm) piston pin location - 903cc - 112.5mm rod length - Rod angle = 1.63
  • 63.5mm stroke - 0.995 inch (25.25mm) piston pin location - 843cc - 115.25mm rod length - Rod angle = 1.81

If we now look at what effect the same changes will have with the "tall deck" blocks, you will see that there is even some advantage for the 903/1050cc motor. These blocks are approx. 5.6mm taller than the short deck blocks. This means that the piston pin will be approx. that same amount lower again in the piston. Note: The 843cc motor combination would not be advisable !!!

  • 74mm stroke - 1.219 inch (30.93mm) piston pin location - 982/1050cc - Rod angle = 1.486
  • 69mm stroke- 1.318 inch (33.45mm) piston pin location - 903cc - Rod Angle = 1.594
  • 63.5mm stroke - 1.426 inch (36.19mm) piston pin location - 843cc - Rod Angle = 1.732

Now, if we again modify the rod length to maintain a high pin position, the following results are obtained

  • 74mm stroke - 0.995 inch (25.25mm) piston pin location - 982cc - 115.7mm rod length - Rod angle = 1.5636
  • 69mm stroke - 0.995 inch (25.25mm) piston pin location - 903cc - 118.2mm rod length - Rod angle = 1.713
  • 63.5mm stroke - 0.995 inch (25.25mm) piston pin location - 843cc - 120.9mm rod length - Rod angle = 1.90

In all cases the rod angularity is markedly reduced, which should result in some reduction of parasitic friction losses.

Quite obviously, in order to make these changes, custom connecting rods have to be used. These must of be the best possible materials and must use the highest grade connecting rod bolts, if the engines are going to survive 8000+ RPM limits. If a good quality 69mm crankshaft were available, then an interesting combination would be a 67.6mm bore with a 69mm stroke using a 118.2mm rod. This would produce a displacement of 990cc.

In an effort to "standardize" rod length configurations and to provide some economies of scale as far as manufacturing is concerned, Scuderia Topolino will make pistons for two rod lengths available, standard (110mm) and overlength (117mm). The Scuderia Topolino General Catalog lists all of the piston/rod combinations available as standard items. Of course if you need a connecting rod/piston combination that is different, please do not hesitate to ask.

CAUTION - Some of the rods will get quite long and only the absolute best of materials should be used. In additional the bottom of the cylinder will likely require notching on most of these combinations and certainly on ones using a short deck block (600/850/1000OT) with a long rod.

All rods supplied by Scuderia Topolino are equipped with ARP2000 or ARPL19 connecting rod bolts. The A112 standard rod is the heaviest at 443 gram. By comparison, an Abarth 1000SP connecting rod (also used in the TCR motors) is lighter at 365 gram. Even after some grinding and polishing, it is still the heaviest. For its size, the 1000SP rod is surprizingly light. Its design is I-beam, only much stouter. The H-beam rods from Carillo weigh in at 384 gram. The Scuderia Topolino connecting rod, with a combination of H and I beam characteristics weighs 350 grams in standard form and less than 330 grams in the “narrow” version.* Finally, you could decide for a titanium connecting rod and this would weight approx. 30% less than the steel equivalent or 220 grams.

3.2 Pistons

Piston metallurgy. There are basically four types of forged pistons on the market today. Three of these are made of an Aluminum/Silicon Alloy, whereas the forth is made of 2618 Aluminium.

  • 2618 Aluminum - No Silicon
  • Hypoeutectitic - Less than 12% Silicon (typically 9 %)
  • Eutectic - 12% Silicon
  • Hypereutectic - More than 12% Silicon

Basically, silicon adds "wearability" to the alloy and in street applications would be a great plus. One drawback is that it makes the aluminium somewhat brittle. For low RPM motors this is not a problem.

For high performance applications, almost all pistons, including those supplied by Scuderia Topolino, are made from 2618 Aluminium or similar, but without silicon. This means that they will expand somewhat more than an aluminium/silicon alloy piston, but they will be significantly stronger.

While I have not specifically addressed cast pistons, the same considerations would apply, plus the cast piston will have a slightly less dense structure and therefore may not be as strong. Although some would claim that the expansion characteristics of the cast aluminum material are also less than a similar sized forged piston, for ultimate strength a forged piston is recommended

Ring Configurations - All Scuderia Topolino pistons are manufactured with the following ring pack. We find that it provides a good combination of wall tension, seating, wear resistance and cost.

  • Top Ring - 1.2 mm Chrome
  • 2nd ring - 1.2mm Nodular Iron
  • Oil control ring - 2.8mm multiple segment oil control ring.

Note: All pistons produced after Feb 2007 will also come standard with "peripheral gas porting" to aid on top ring sealing.

It is possible to use a 2-ring configuration. This would be purely a "competition only" setup. The overall size of the ring pack would be reduced by about 3mm, allowing the piston pin to be moved up an equal amount. This would be particularly helpful for 74mm stroke motors. The following results would be obtained.

  • 843 block motor - 74mm stroke 113mm long rod for a rod angle of 1.527
  • 903/1050 block motor - 74mm stroke 121mm long rod for a rod angle of 1.64

See explanation of rod angle issues under the connecting rod section.

Piston Layout Combinations Scuderia Topolino pistons are now in their 3rd design generation. The current piston has a reduced skirt length, thinner skirt sections, tapered forged piston pin, and the top ring groove is now gas-ported for improved top ring sealing. Weights of all of these new pistons have been reduced by an average of 15%. As an example, the 3rd generation piston with pin now weights 218 gram (compared to 279 gram for the 2nd generation piston).

Scuderia Topolino now stock three different piston dome configurations. There are:


1. Flat Top. These pistons have approx. 10.5:1 compression with a standard Fiat/A112 cylinder head. These is perhaps 0.5 point higher compression than the standard cast A112 piston, which is slightly dished.


2. Small Dome with Valve Reliefs. These pistons have approx. 12:1 compression with a standard Fiat/A112 cylinder head.


3. Large Dome with NO Valve Reliefs. These pistons have more than 13:1 compression with a standard Fiat/A112 cylinder head. Caution: Depending on the type of camshaft used, these pistons may produce more higher dynamic compression than desired.

I am sure that someone will ask why we did not make it a "full slipper style" piston. The overall goal was to reduce the weight of the pistons with pins to 250 grams or less. We have more than achieved this without resorting to a full slipper design. These were the considerations.

1. Because of the small diameter of the piston, in order to achieve a full slipper design, the engineers felt that the pin bosses would become too small to support the RPM levels expected from the pistons. At 8500 RPM the piston speed is around 68 ft/sec (21 m/sec), with 3900 G’s of force at TDC at the same RPM.

2. It was determined that not much additional weight advantage could be gained from a full slipper design, as the amount of aluminum material saved by repositioning the pin webs would me minimal.

3. The piston would retain better dimensional stability with a semi-slipper design with more predictable expansion characteristics. This would promote better ring stability and cylinder sealing.

Here are the weight comparisons between 2nd and 3rd generation pistons.

Abarth/Fiat 10.5:1, 3.5mm dome with valve pockets - 2nd Gen. 279 gram - 3rd Gen. 218 gram

Abarth/Fiat 13:1, 6mm dome, no valve pockets - 2nd Gen. 275 gram - 3rd Gen. 226 gram

Abarth TCR 10.5:1, 2nd Gen. 264 gram - 3rd Gen. 239 gram

Abarth TCR 12.5:1, 2nd Gen. 274 gram - 3rd Gen. 249 gram

Anti-Friction Coatings - Developments in anti-friction coatings have come a long way in just a few years. The coatings available today certainly provide a measurable advantage when used in a race motor. At the moment I specify coating applied by Calico Coatings. This plasma sprayed coating can be used on piston skirts, main and rod bearings, rocker shafts, lifters etc. They also have oil shedding coatings for use on reciprocating parts such as crankshafts and connecting rods.

In January 2007 we started testing with DLC type coating by Sorevi (Bakaert). This company has developed several proprietary coatings by the trade name CAVIDUR. These can be applied to various engine "friction interfaces" such as rocker shafts, camshafts, lifters, finger followers and quite interestingly pistons skirts. I have also done a complete gearbox to see how the material stands up high levels of both friction and shock loading. Certainly on rotating shafts, collars and plain bearings, the material would appear to have distinct anti-friction advantages. We examined a set of lifters, camshaft, pistons and piston pins, after a full season of races and found virtually nil wear to these items.

Flame Propagation – In order to get the maximum combustion pressure from the engine, measures must be taken to insure that there is complete burning of fuel. Just because an engine has 13.5:1 computed compression does note mean that if will produce the most power, unless it can burn the fuel efficiently. A good general rule of thumb is that the higher the dome on the piston, the more difficult it is going to be to burn all of the fuel in each combustion event. In essence, the fuel charge will be distributed on either side of the dome and the spark plug is on one side of the dome. Therefore an engine running with 12.5:1 compression, and completely burning all of the fuel in the cylinder in each combustion event, may in fact produce better horsepower. About the only way to tell is to run two engines, back to back, on an accurate engine dynamometer.

3.3 Balancing Considerations

Rotational Assembly Imbalance Tolerance - and its affect on engine reliability

No one involved in the preparation and building of race engines would argue the importance of balancing. Improper, or better yet inaccurate, balancing may have far reaching implications on both the performance and the useful life expectancy of the engine in question. Balance, or the lack of it, is more than simply a matter of matching individual components. Perhaps more importantly, it is the balance tolerance limit of the entire assembly that is of vital importance. As I will explain later, a small static weight difference can have a profound negative effect at higher RPMs.

In dealing with historic engines, like the Fiat 600 engine and it many derivatives, one has to come to grips with the fact that this engine was designed well over 50 years ago, and probably to a much different design and performance criteria than used in historic motor sport competition today. After all, it is a fairly plain, garden variety 3 main bearing, OHV motor. For this very reason balance tolerance limits may be much more critical here than in say a modern, five main bearing engine such as the Duratec 2.0 litre 4-cylinder. Yet amazingly, at high levels of tune, it is not impossible to get power outputs of 100 HP/Litre from these early Fiat block designs. Over the years I have seen Fiat blocks fail in two general ways that could be attributed to "excessive balance tolerance".

  • Center Main Bearing Support Failure - Here the center main bearing portion of the cast iron block literally breaks away from the block. It is sometimes difficult to determine cause/effect with this failure if it is determined that the block is broken after the engine has had a major mechanical failure (thrown connecting rod). One might argue that the rod broke first, and then the block was damaged when the block was struck by the rod. I believe the cause/effect sequence may be different. I believe that due to excessive imbalance tolerance the center main was pulled from the block and then due to crankshaft flexing further damage is inevitable.
  • Cracked Front Block Surface - The failure I have seen on several A112 blocks where there is a fracture between the front main bearing bore and the front cam bearing bore. I believe this to be entirely an excessive balance tolerance problem

So, what is involved in achieving a minimum balance tolerance. First we have to understand the nature of imbalance, and how the affects of this imbalance are manifested, and then work backwards from there. Imbalance is generally caused by some type of "uneven distribution of weight". In the case of a race engine, imbalance has to deal with rotational as well as reciprocating elements. This uneven distribution of weight can be cause by several factors including

  • Improper manufacturing and installation tolerances
  • Metal inconsistencies (forgings or castings)
  • Fasteners
  • Trapped oil
  • Overall component and assembly weight
  • Damping (or lack thereof)

First it is important to understand the balancing procedure. Almost all dynamic balancing is done at rotational speeds between 200 and 1200 RPM. This is far from the 7000-9000 RPM that these engines are likely to see in competition. Therefore, any small imbalance at 1200 RPM will be much more pronounced at 9000 RPM.

Second, many "assume" that the crankshaft, the major engine rotational component, is stiff enough to resist bending caused by compression loads imposed on it, given that it is adequately supported in a crankcase of sufficient "beam stiffness". These are very large assumptions when it comes the 3-main bearing Fiat blocks, so it may be prudent to keep the balance tolerance as low as possible.

Let’s start by defining the types of mechanical issues that can affect balance. Most engine builders would concern themselves with the rotational balancing of the crankshaft. However, this process is more complicated that first meets the eye. The crankshaft is made either from cast iron, nodular iron, steel casting, forging or billet, in corresponding order of resistance to bending. Given the irregular shape of the crankshaft, as well as the other assembled components that make up a complete crankshaft assembly, the job of balancing is made all the more difficult. Certainly fully machined forgings or billets are the most preferable, as they should have equal dimensions for the various webs and counterweights associated with the crankshaft itself. Cast crankshafts, other than being more flexible than their forged or billet steel counterparts, may also have casting inconsistencies, causing differing dimensional characteristics and associated imbalance. Certainly it is helpful if these rotational dimensions are standardized, before any balancing is done.

Illustration attributed to Steve Smith from Racecar Engineering article "Vibration Free"

It is also important to inspect all of the ancillary attachments to the crankshaft to make sure that they are concentric. The tin front pulley of the Fiat 600, if bent or out of round should be discarded, as it may have both parallel and/or angular displacement, contributing to overall crankshaft assembly imbalance. Better yet, it should be replaced with either a light alloy or steel machined pulley where the outside diameter is concentric with the central bore, and the front surface of the pulley at 90 degree to the bore centerline. With the pulley centerline the same as the crankshaft snout centerline, this will insure the minimum of parallel and angular displacement. The keyway bore should not have any excessive wear or clearance, as this will allow the pulley to rotate in relation to the keyway. The same argument would also hold true for the flywheel and pressure plate assembly.

There is an "order" of balancing that I recommend, but first we need to look at what type of imbalance tolerance we are willing to accept. It is simply not enough to tell the machine shop to "balance the assembly".

We can illustrate this by first balancing the four associated connecting rods (make sure the bearings are installed and rod bolts seated). This requires a scale, accurate to 0.1 gram or better, and a connecting rod weighing fixture capable of supporting the small end of the connecting rod so that the rod centerline is level and square to the top of the scale platform. The large end of the rod should be placed in the center of the scale platform, and as the rod is a fixed length, the small end will be the same distance away. In this way we can match the connecting rods so that the big end weights are within 0.1 gram of each other. How important is this measurement? According to Steve Smith, of Vibration Free" in England, "If one rod was 0.1 gram heavier than the other three, this would produce a force of 5.5 lbs (2.5 kg) at 6000 RPM for an engine with a 70mm stroke". As Abarth motors of 982cc displacement have an even large stroke (74mm), and will achieve engine revolutions of up to 9000 RPM, for a 0.1 gram imbalance the amount of force would be more than 13 lbs (5.9 kg). An assembly that was out by 5 grams would generate a force of 659 lbs (299 kg) at 9000 RPM.

Engine designers calculate the engine bearing surface to be able to support a certain load force at a given oil pressure. If this "loading limit" is exceeded, then meta-to-metal contact occurs and a catastrophic failure is inevitable!!

Obviously this is just one component out of many that may affect the balance of a complete crankshaft assembly. Once the big ends of the rods have been balanced (probably to 0.1gram or less) then the overall weight of the rods must be measured and recorded. It is helpful to temporarily number the four rods with a marking pen. Next weigh each piston assembly (complete with rings pins and clips) at match them to the respective rods so that the combinations are as close to the same as possible. Only them can you remove weight from the small end of the rods that are heavy, as compared to the lightest combination of rod/piston. Once you have them all the same, then you can permanently mark the rod/piston combinations for the cylinder where they will be installed.

Every component that rotates or oscillates has an inbuilt resonant, or critical frequency. Crankshafts may in fact operate at higher critical speed than their natural frequency and will need to be balanced in several planes along their length. As any imbalance is amplified at/near the critical speed of any component, balancing at multiple planes is essential. It is this amplified imbalance at different locations along the crankshaft that causes bending and flex, resulting increased parasitic losses and if too great engine failure.

The order of things should be as follows:

  • Crankshaft only
  • Add front pulley and nut
  • Add flywheel
  • Add pressure plate

Now you can replace any one of the ancilliary items attached to the crankshaft without having to rebalance the crankshaft itself.

There is one more factor that can be lessened, in terms of imbalance tolerance, and that is the effect of any oil clinging to the rotating and reciprocating assemblies. Engines running at high RPMs create what an "internal whirlwind" of air and oil, around the crankshaft. This is referred as "oil roping". This may add to any other imbalance that exists. This can be minimized by the following methods.

  1. Installing a windage tray - This is a full separation between the rotating crankshaft and the oil in the pan. This will keep the oil from contacting the crankshaft during cornering, acceleration and deceleration.
  2. Are oil pan windage trays important in preventing oil aeration and why? The answer is both yes and no. All engine oils have a level of dissipated air, as part of their makeup. At one atmosphere (14.7PSI) it is generally accepted that there is 9% by volume of dissolved air in mineral oil (Bunsen's Coefficient). Therehave been several papers written about the behavior of oil within the sump of a wet-sump lubricated engine (dry sump engines have a different set of circumstances). From the various studies, on the effect of windage tray design on surface aeration of oil, it would appear that the even at 50 PSI (3.4 times higher than atmospheric) the percentage of allowable air entrainment in oil may be on the order of 50% for rotating assemblies such as crankshafts, camshafts and counterbalance shafts. That is not to say that 50% should be the design goal. As the percentage of entrained air in oil tracks both RPM and oil temperature, it is important to control free air in oil to a workable amount, as concentration higher than 50% will cause main failures and concentrations of 30% is considered by many to be the upper limit for connecting rod bearings due to interrupted nature of the oil feed.

    From empirical testing at Ford Motor Co. the effects of oil droplets, flung from the rotating assemblies, on the free air percentage of oil in the sump (prior to entering the pump) is inconclusive. Yes, at higher RPM these high speed droplets did cause the oil to foam on the surface of the oil in the sump, but this appeared to have little effect on the entrained air percentage of oil entering the pump. Adding a windage tray to a wet sump appeared, according to cited evidence, to have little positive effect. Once oil enters the pump no additional air is added to the oil as it traverses the system. With changes in pressure there may be a conversion of some free air to dissolved air and vice-versa. A windage tray, or some other means of controlling the movement of the mass of oil in the pan, does have considerable positive value in terms of ensuring that the oil pump pickup does not become uncovered during high G force situation (Braking, cornering and acceleration - in order of magnitude), or preventing the crankshaft from causing additional aeration from physical contact with the oil in the sump.

    So it would appear that the principle benefit of a windage tray is to control the "gross" movement of oil in the sump. However, while we are examining this situation, lets take it one step further.

    Because there is both dissolved and free air in the oil, the two must be considered together. Most test specifications dealing with oil aeration assume an 18% air-to-oil volume for dissolved and free air combined. This number also appears to be a good average to shoot for in an oil system design. As I understand it, using Bunsen's coefficient for engine oil, "for every one Bar (14.4 PSI) increase of oil pressure, the oil can take up an extra volume of air equal to 9% of the oil volume". As such at 3.5 Bar (50 PSI) the volume of air-to-oil could be as high as 31.5%. Using the old "hot-rod" formula of 10 PSI oil pressure per 1000 RPM the allowable percentage of air-to oil would equate to 49.6%.

    So is this high percentage of air-to-oil what we want? The answer is patently NO. Quite the opposite is what is required, as there is a mechanism within the oil delivery circuit that will negatively impact the percentage of free air in oil to certain parts of the engine, particularly the upper end. Generally oil routed to the cylinder head will travel via a "sharp edged restrictor". The effect of this type of restrictor is to cause an oil pressure drop in the oil traveling to the upper end. Typically, the pressure drop will be in the order of 40%. Thus if the pressure at the main bearings were 3.5 bar, then the pressure at the rocker shaft (in the case of the Abarth motors that I work with) would be on the order of 2.1 Bar. Whereas at the oil pump, due to the increase in pressure over atmospheric, the percentage of allowable entrained "could" be more due to the higher pressure, this might place the upper end lubrication at risk. The lowering of pressure at the orifice feeding the upper end means that dissolved air may reverse from dissolved to free air. Combine the additional free air with reduced oil flow (a function of pressure and orifice size), then it is quite feasible for rocker/shaft boundary interface in an OHV engine design to be marginalized. In competition engines with more aggressive camshafts and higher rate valve springs this would give serious cause for further analysis.

    One could argue that it would be beneficial to reduce the size of the restriction, thus increasing the pressure/flow to the upper end of the motor. This would have a beneficial effect on reducing the pressure drop, thus lowering the possibility of reconstituting free air from dissolved air in the oil. In addition, the additional oil to the upper end would enhance cooling due to the increased flow. However, this may decrease overall engine oil pressure and yet other measures may have to be taken to return upper end oil to the sump in an efficient manner.

    So there is a fine balance to be struck if one were to contemplate running lower oil pressures to reduce parasitic losses. Certainly the use of a windage tray to control gross oil movement in the oil sump and keep the oil from coming in contact with the rotating assembly is a good idea, as the lower the entrained air percentage per volume of oil, the better. Further, the use of an effective "oil scraper" to strip oil from the rotating assembly and return it to the sump would also be recommend.

  3. Installing a crankshaft scraper - A crankshaft scraper is a device attached to the main caps of the engine that is contoured to very close tolerance of the counterweights and cap areas of the connecting rods. The scraper assembly may be as close as 0.015 inch (0..4mm) of these surfaces and is intended to scrape any remaining oil from the crankshaft counterweights and rods.

  4. The combination of an effective windage tray and crankshaft scraper will greatly reduce the impact of oil on the balance of the crankshaft assembly. Perhaps even more importantly, by reducfing the amount of oil clinging to the crankshaft and connecting rods dyno tests have shown an increase in horsepower of up to 5%.

    This is of course for wet sump situations. If a dry sump is used a windage tray is not necessary, but a good crankshaft scraper my still reduce parasitic roping losses.

    The instances of block failures are, in all likelihood, due to larger than allowable imbalance tolerances. A large out-of-balance condition may cause the beam stiffness, available from the standard Fiat block, to be exceeded. This will cause the engine to "pound the bearings", causing eventual main bearing and/ or block failure. Plus, the effect of maintaining strict balance tolerances will pay off in added reliability and performance due to the reduction of parasitic losses.

    My thanks to Steve Smith of Vibration Free in Oxon UK for some of the material used in this discussion.

3.4 - Compression Ratio Fundamentals

There is a basic difference between “Calculated Compression” and “Dynamic Compression”. Both numbers are important, but the one that will make difference in the reliability of a race motor will be the dynamic compression.

Calculated compression is basically the ratio between the volume of the cylinder and combustions chamber, divided by the combustion chamber volume. In the case of a A112 motor with a standard cylinder bore, this could be as follows:

Cylinder volume (per cylinder) 259.75cc

Plus, Combustion chamber volume 28.00cc

Equals 287.75cc

Divided by, Combustion chamber vol. 28.00cc

Equals 10.28

So this motor would have a Calculated Compression of 10.28:1 and would run quite happily on high octane pump fuel.

For most competition engines we will want to raise the compression ratio to around 12.25:1, and in some cases we may even go as high as 13.5:1.

This calculated compression ratio is the beginning of a much more complicated calculation to determine the “knock resistance” of the engine. This will involve both the camshaft intake valve closing point, the octane rating of the fuel to be used and, the Dynamic Compression ratio.

First, we have to compute Dynamic Compression ratio. I am not going to list the actual formula for computing it here, but this information can be found in the internet for those who are really interested.

Here is a comparison of different camshafts in a 1046cc motor with 110mm rods, 12.25:1 static (computed) compression ratio.

Camshaft

Duration/Lift

L/C

Overlap In.

Int. Closing Deg

Dynamic Displ (DD)/DCR

SLR300S

290/300/12.4mm

107 deg

81 deg.

69 deg. ABDC

804cc/9.49:1

SLR300

300/12.4mm

108 deg

84 deg

74 deg. ABDC

765cc/9.07:1

Kent FT 6

304/10.8mm

106 deg

92 deg

78 deg. ABDC

732cc/8.73:1

PBS A8

305/10.6mm

108 deg

89 deg

76.5 deg. ABDC

745cc/8.86:1

CatCams

305/11.45mm

108 deg

89 deg

80.5 deg. ABDC

711cc/8.50:1

CatCams

310/11.45mm

108 deg

94 deg

83 deg. ABDC

689cc/8.27:1

Abarth 316

316/10.4mm

105 deg

96 deg

88 deg. ABDC

644cc/7.80:1

Laur 319

319/10.5mm

108 deg

103 deg

87.5 deg. ABDC

649cc/7.84:1

Abarth 336

336/11.7mm

105 deg

126 deg

93 deg. ABDC

634cc/7.70:1

From the above, it is obvious that if someone asks you what compression ratio you are running, the answer will be meaningless, without telling the person asking the question, additional details.

The formula for computing horsepower is: Horsepower = rpm x torque / 5252

From this you can deduce that if all of these camshaft combinations were to produce the same “peak” horsepower, then those camshafts with smaller valve overlaps, and early intake valve closings, will produce higher torque/horsepower at lower RPM and taper off before 8000+ RPM, whereas the cams with large valve overlaps, and later intake valve closings, will produce lower torque numbers and will rely on higher RPM levels (perhaps as high as 9000+) to make the same “peak power”.

Peak horsepower is great for bragging rights, but we don’t actually spend a great percentage of our time running at 8000 RPM or more. Much more time is spent between 5500 and 7500 RPM, so this is where we should aim for the best engine efficiency. The better indication is to plot the torque/horsepower numbers 5000-7500 RPM and to find the highest “average” horsepower and torque in that range. Then choose a camshaft that will best deliver this. This will, in most cases, provide the best overall performance in a road racing vehicle.

There is a direct relationship between the "Dynamic Displacement (DD)" of an engine and the closing degree of the intake valve. This, then in turn, determines the DCR of a motor. This is one of the critical design criteria in building any race motor. The earlier that you can close the valve, without creating negative pumping effect, the greater the DD and the resultant DCR. You can also vary the DCR by advancing/retarding the camshaft up to 4 degrees. The higher the DCR, the greater the knock sensitivity, and therefore the greater the fuel octane requirement.

You will notice that this one engine can in fact produce a Dynamic Compression (last column in the chart) ranging from 7.70:1 to 9.49:1. It may not seem immediately obvious, but it is unlikely that all of these engines will survive using fuel of the same octane rating. From my own experience I know that a dynamic compression of 7.99:1 will have a “knock rating” of approx. 4.4 @ 6000 RPM. This is very close to being marginal for 100 octane fuel, but with no more than 28 deg. of total distributor advance. However, changing the camshaft to one that produces a dynamic compression of 8.99 will require a fuel with an octane rating of 105 in order to maintain a knock rating of 4.4 or less. Now if we jump up to the most aggressive camshaft, with a dynamic compression of 9.45:1, nothing less than 112 octane racing fuel will do.

As you have probably noticed by now, compression ratio cannot be viewed alone, without taking into account ignition timing and fuel octane. I suggest that all who are interested knowing more about "fuel octane" numbers go to the following link and read the information carefully. Here in 3 pages is a very good explanation of the different rating measurements and how to compare different rating systems.

http://www.btinternet.com/~madmole/Reference/RONMONPON.html

After you have read and understood this information, the following will make more sense.

Note: I am running my own engine at a computed compression of 13.5:1, The camshaft that I an using computes to a DCR 10:57:1. With DCR levels this high you MUST use very high octane. For this motor I use Sunoco 112 octane leaded racing petrol. In the interest of "longevity" I have reduced the static compression ratio to 12.5:1, which has lowered the DCR to approx. 9.67:1

 

 
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