Old connecting rods...
 

Hi,
Trying to get a feeling for the viability of re-using original connecting rods.
Not a 911 guy, but all the concepts still apply to the water-cooled world...

You guys seem pretty content revving the nuts off your engines.
Some of you have con rods that are approaching (or even past!) 50 years old.

I know that the fancier builds get aftermarket rods but what are the real limits of the stock pieces, if you keep the revs down? Is age a factor?

Anything to consider besides new bolts, re-sizing the big ends and pin bushings? You guys shot-peening the rods to stress-relieve the surfaces? Anything more?

The ads which state 'shot peened to relieve stresses' clearly demonstrate a lack of understanding of this process and I think originates from China.

Shot peening is a process which, if carried out correctly, can significantly improve the fatigue performance of many metallic components

By bombarding the surface of a metal with a smooth round shot in a controlled manner it is possible to introduce a significant level of compressive stress and this is extremely beneficial in terms of fatigue endurance.

Fatigue is responsible for around 95% of mechanical failures and at a basic level is easy to describe.

In simple terms most fatigue failures can be described as a two-stage process, Stage 1 is fatigue crack initiation and Stage 2 is fatigue crack growth.

Fatigue crack initiation generally accounts for 90% of the life of a component.

Fatigue cracks normally develop on the surface of a component and initially form due to what is referred to as 'slip band intrusion and extrusion' which are crystallographic defects that are virtually impossible to detect.

The development of these defects always require the presence of traction vectors which are usually caused by shear stresses and commonly form in specific crystallographic directions.

The presence of residual compressive stresses means that the surface tensile stresses must be of a much greater magnitude to cause the damage that would occur if the surface were unstressed.

Nitriding has a similar influence on fatigue life as it also introduces surface compressive stresses but Shot Peening avoids the need for heat treatment.

If is carried out correctly shot peening is very beneficial to components such as Turbine Blades, Con Rods and many other critical parts. If it is carried out badly it can have a negative impact on fatigue life and can introduce defects.

The quality of the shot, the screening of shot to remove broken particles as well as the velocity and the direction of application all have a significant influence on performance.

With used components it is difficult to make a decision with regard to the amount of fatigue damage that has been accumulated unless the part has been used in very controlled conditions and monitored in terms of the stresses that have been generated in service. Many Aircraft components are monitored in this manner but the costs of producing the base line data to make this approach work is very high.

Once a detectable defect has developed any component affected should be discarded.
With a used con rod it is possible that the part could have used the majority of its 'fatigue initiation' life and may be ready to crack.

By careful shot peening it is possible to remove the accumulated damage and effectively re-life the part.

By using companies that are FAA/CAA Approved it is very likely that excellent results will be obtained at a very reasonable cost.

I would suggest that Con Rods are shot peened in the following manner.

Shot Peen 200% Coverage 6-10 ALMEN "A" WITH #230 Austenitic Stainless Steel Shot.

Would there be any benefit to cyro treat them as well? What is best, cryo then shot peen or vice versa?

Cryo Treatments are interesting and IMHO they should be considered with a great deal of cynicism.

There is no doubt that many tool steels, particularly those used for the manufacture of high precision parts benefit from this type of treatment as it generally assist with dimensional stability.

The main benefit of this process is for the treatment of steels that suffer with problems of retained Austenite following heat treatment.

Retained Austenite is a metastable phase and there is a thermodynamic driving force which wants to allow its transformation to a Martensitic type of structure. The activation energy to allow this transformation can be supplied from either a mechanical force or can be thermally activated either by low temperatures of increased temperatures.

This transformation involves a small expansion in the volume of the component and this can lead to distortion and subtle changes in shape.

When considering close fitting parts or transducer materials for the manufacture of load cells these very small variations in size can cause problems and it has been common practice to cryogenically treat parts prior to final grinding.

Once the steel has been cooled to typically Liquid N2 temperatures full transformation will occur and the parts will be stable.

The type of material used for the manufacture of con rod, which would typically be a 4340 type steel and normally these steels are fairly stable due to the influence of carbon and tempering effects which make the retained Austenite relatively stable and immune to the effects of mechanical stresses. The volume of retained Austenite in 4340 type steels is typically in the region of 2% and tempering above 600 degF will cause it fully transform.

In terms of metallurgy there is no benefit to be obtained in cryo treating this type of steel in terms of either mechanical properties or dimensional stability.

There has recently been some evidence to suggest that the cryo treatment of chrome-silicon wire springs can result in an improvement in fatigue performance compared to an untreated spring and this is due to the development of compressive residual stresses at the surface of the wire.

The basic measurements of the residual stresses suggest that this influence is less than would be expected in a shot-peened spring but no comparative data has been produced.

Frankly I wouldn't bother but it will do no harm and if you are going to cryo treat them do it first.

I carried out a significant amount of work on the fatigue behaviour of shot peened gears more years ago than I like to admit and presented a paper on this subject in Amsterdam and have followed this subject with interest for many years.

Some work more recently published in Sweden on behalf of Scania reported the following observations:

In order to find out how the fatigue limit is affected by the material
hardness, a number of gears were tempered whereas other remained untempered.
Similarly the retained austenite content was altered in some of the gears with a
cryotreatment.

The shot peening of these gears was performed by dual shot peening (200% Coverage)

The results show that the fatigue limit is enhanced when the shot peening was
performed with hard media.

The amount of retained austenite does not seem to affect the fatigue limit for samples shot peened with this hard media.

Ollies Machine Shop, inspected for straightness, resized for ARP torque spec, re-bushed the wrist pin holes and Honed both.

I feel good about my 32 yo Rods :D, the New ARP kit will keep her together just fine.

I'm no expert, but considering chris _ seven' s response. I would take his advice. But then, if one has a " budget " .. I know most don't think we do with our hobby and will simply use it as a daily driver. Then rebuilding with new rod bolts like ARP is good. Otherwise, start fresh with some of the higher price point rods from the start. ..how big is your budget for your hobby?

The rods in a 911 are HUGE and ROBUST.

I think these things are barely stressed given their mass and being really, really short.

Chris: "With used components it is not (?) difficult to make a decision with regard to the amount of fatigue damage that has been accumulated unless the part has been used in very controlled conditions and monitored in terms of the stresses that have been generated in service. Many Aircraft components are monitored in this manner but the costs of producing the base line data to make this approach work is very high."

This is the crux of the matter, right? I wonder about "not"? As in, it's hard to determine whether to reuse a component unless the stress has been tracked in some manner? It seems you're discussing damage at a much smaller level than can be measured especially by DIYer's. Are there any tests that don't involve a lab to see stage 2 damage, something like a more sensitive magnaflux of some sort? Or, in your opinion, does Tippy have a point and that, say, a "T' engine's rods most probably won't have seen stresses that rise to stage 1/state 2 damage because the part is over-engineered for lower HP engines?

Here is an aftermarket 8000 RPM capable big-block Chevrolet connecting rod used in a 7.4 L (454 ci) engine that uses a MONSTROUS 101.60 mm (4 in.) stroke:


http://forums.pelicanparts.com/uploa...1489161179.jpg

Now, here are our beloved, stumpy little girthy connecting rods used in the, don't exceed 7000 RPM (in the bigger 911 motors it seems the concensus is) engines at a teeny stroke of say, only 74.4mm for a 3.2 (of course the smaller engines have smaller strokes yet down to 66mm!!!!):


http://forums.pelicanparts.com/uploa...1489161365.jpg

I just can't see how we are stressing these things IMO?. Maybe I am missing something, but a stock 911 rod is a beastly little thing!

Well that was a 'dumb' typo on my part - Thank you for pointing out I have now edited this mistake.

The problem with metal fatigue is that up to the point at which the crack physically initiates there is no real way to test the part to understand how much damage has accumulated.

The rate at which cracks propogate once they have initiated is very important and the conservative decision is always to assume that crack initiation is effectively the failure point.

With some components it is possible to develop a test programme to decide when failure will occur and simply replace the part at some safe proportion of a components life.

Ti rods used to be systematically replaced after around 50 racing hours and many military aircraft parts used to be treated in a similar manner.

This is because for many components the life between crack initiation and failure is very short and the risk of failure is too great to tolerate.

If parts can support reasonably long cracks it is possible to use an inspection based technique as long as the crack length does to reach a critical length between inspection periods. In this way if parts were crack detected and found to be defect free they could be re-used.

I would generally agree that 911 rods are conservatively designed and under most circumstances will never suffer from fatigue failures.

The decision is a simple risk/reward analysis.

If I were building a 2.0 litre engine to run at 8500rpm I would almost certainly shot peen the rods.

The cost is small compared to the damage that could occur if a rod failed.

If I rebuilt a standard 911T engine I wouldn't generally bother.

That's been my point. I think nearly all rods out the side of a case were from oil starvation. :)

Stock Porsche rods, with the exception of the 993 ones, are plenty strong and do not fail unless the bearing has an issue.

The main problem with OEM rods is they are heavy and that exponentially loads the bearings above 7K RPM. Porsche did use a nitrided OEM rod in the 2.8 & 3.0 RSR engines, but that was due to FIA Gp 3 homologation. The factory usually used a Ti rod in almost all race engines.

We prefer Pauter or Carrillo 4340 rods in all of our race engines for weight reduction and additional strength. (Arrow also makes an excellent product). Special apps get Ti rods, however these are life-limited so they are not for anyone on a budget.

Has anyone actually seen a Porsche Titanium rod have a fatigue failure? I know PMNA doesn't always replace them in 3.6L Cup motor rebuilds and with people running 996 Cup motors 200+ hours without rod failure and GT3s accumulating 100+k miles (slightly different rod to Cups but still Ti) with the rods intact, are we sure that they are actually operating beyond the fatigue limit?

What signs does the rod bearings exhibit?

Porsche Titanium rods are usually replaced by PMNA at 60 hours yet several set are running at 100+ i and the street version of the GT3 is not pulled into the Dealer at any preconceived time and replaced Why?

Some great advice given. Like all advice, it is only as good as the person receiving it wants to use it.

If you are talking about 911/930 Rods, I have never seen one fail if it was used as designed. Same for 944 Turbo Rods.

If they are rebuild correctly with care, they will do a great job. Its my opinion that all too often aftermarket rods are used when not required. I see Carrillo H section rods with Carr bolts in engines that make 250 HP and run at factory RPM limits.

Any part of an engine should be used based upon the engines use, performance and the parts used if its design falls within the engine use criteria. If there is a fitment factor, pump clearance, length or other design features that are required, then an aftermarket rod has to be considered. Weight is another factor. However, many aftermarket rods are as heavy as factory Rods.

If you are wondering about re using 944 Turbo Rods, have them rebuilt correctly, including all typical checks, handle them with care, and as suggested have them shot peened.

I am amazed at the mishandling of engine parts, I often see. Parts thrown down on the bench, put inside boxes without separations etc. This is one of my pet peeves, as is calling air cooled cylinders "jugs"!You put milk in jugs and pistons in cylinders!!

Just make sure that you do not induce the stress areas with nicks etc by mishandling. The engine will make you pay for it. I can assure you.

If the rods aren't breaking they simply haven't accumulated sufficient damage to cause failure, the issue is when will this happen? :)

This is a very difficult question to answer and I would reiterate that a decision to use a part beyond the life recommended by the manufacturer is a simple risk/reward judgement.

As we have no data in terms of levels of stress and no real fatigue test results you simply pay your money and make your choice.

The entire problem is based on the stochastic nature of the fatigue process, the variability of the material we are considering and the loading spectrum that creates the stresses that drive the fatigue process.

The concept of a Fatigue Endurance Limit is interesting and has been well accepted for many years and in practical terms seems to work - at least for the majority of conventional engineering steels.

The main reason for the development of a Fatigue Endurance Limit is concerned with a metallurgical process known as strain aging. In this process the dislocations present in the material's crystal structure are pinned into place by solute atoms which produces a very localised increase in yield stress and this effect is activated by the cyclic loading that causes fatigue. By using Electron Microscopy the development of dislocation 'tangles' around these solute atoms can be observed and this would suggest that this explanation is valid.

Engineering Steels exhibit high levels of strain aging but this behaviour is not so pronounced in 6AL4V.

It is, however, fair to say that this concept is routinely questioned and recent work published in Acta Metallurgia suggests that when we consider Very High Cycle Fatigue some of the assumptions don't add up. (Cycles of 10exp10 to 10exp12)

Some recent work carried out at UCLA on Ti Turbine Blades subjected to very high levels of vibration suggest that some of the previously applied properties may have been overestimated by at least a factor of 2 .

It is also fair to say that we are unlikely to see this number of cycles within a typical car engine.

Traditionally Ti Alloys were not considered to have a Fatigue Endurance Limit but
it has become customary for the manufacturers of Ti alloys to now quote Fatigue Endurance Limits of around 500MPa at 10exp8 cycles for 6AL4V but I believe if you design components with this level of gross stress you will suffer significant numbers of component failures.

If you read the 'small print' you will see that the 500MPa is an 'estimate' based on 50% of the yield strength of the material and that 'service and geometric' factors must be considered and that this figure should be down-rated between 1.5 and 4 times depending on the specific design.

Very helpful !

I started fatigue testing 6AL4V Ti on a commercial basis many years ago as it was essential to fatigue qualify material used in the manufacture of Sea Harrier components on a batch to batch basis due to material variability.

The thermo-mechanical processing of 6AL4V introduced significant variability which resulted in around 30% or the material supplied being inadequate for the duty cycle

I am sure that modern process control has improved this situation significantly and that by using FEA tools it is relatively straightforward to understand the stress/strain conditions applied to the rod and again using good quality published fatigue data it is possible to design a satisfactory con rod.

The reasons for the 'short' life is most likely due to the inevitable uncertainties and to completely eliminate the risk of a failure.

It is also worth pointing out that improvements in the Fatigue Limit of 6AL4V of up to 10% have been reported by using controlled shot peening and I would consider this to be a worthwhile cost on high value race engines.

Metal Improvements in the UK (part of Curtiss-Wright) have a CAA/FAA approved method for refurbishing Ti helicopter Rotor parts that has been backed up with fatigue test data.

It is also worth noting that it has become a common practice to coat Ti Rods using a Chromium Nitride (CrN) applied using a PVD chamber as an alternative to the plasma moly that used to be applied to the side faces of the big end.

I am not sure I like this process.

Some recent work has been carried out by Bodycote who have investigated the effect of PVD coatings on the fatigue life of 6AL4V Ti landing gear components.

They wanted to apply these coatings to reduce wear and eliminate galling. They evaluated Titanium Nitride, Chromium Nitride and a DLC based coating.

In all cases these coating reduced the axial fatigue life of the components being evaluated and on this basis I would need to see some data before I used this process.

Ti rods will of also be more prone to failure if they are nicked or marked as they are likely to be less damage tolerant than a steel such as 4340.

One day there will be connecting rods in carbon, Lamborghini studies that of meadows! :)

Chris, I got to work with a program off and on for a jet engine manufacturer where the engines were overflown beyond their original intended life cycles. Sometimes, cracked parts were purposely put into the test engine to see if catastrophic failure would occur. Pretty cool project looking back. We also had to gently tear the motor down, but do not touch anything. It was all about allowing the engineers to look at witness marks and possible gasket ot hardware install anamolies. I remember over the course of 15-16 years of working there, engine life cycles were increased by a healthy margin by the engineering group.

Guess my point and you seem to post above, maybe there hasn't been enough testing to fully understand Ti's true life?

I mean, who wouldn't error on the safe side? There's a lot to lose if you get it wrong.

I am sure this will happen but ultimately I think that a nano-platelet Graphene reinforced material will be the final solution and probably use a powder metallurgical manufacturing route.

Recent developments in Graphene manufacture using Soya Beans will impact on the cost of manufacture of the World's strongest material and there are many exciting applications just around the corner.

I hope that this type of connecting rod will be for sale one day for our toys, we do not stop the progress!

If someone would post the mass of a complete piston assembly (rings, wrist pin and circlips included) plus the mass of the connecting rod, I'll run the calculation for the max tensile and compressive force applied to the connecting rod at a given rpm and bmep. Then we will quantitatively know what alternating force we're dealing with.

I'll also need the mass of a complete connecting rod. Knowing this, I'll apply the 2/3 - 1/3 mass distribution convention...that is 2/3rd of the mass is at the big end, 1/3 mass is at the top end. The top end mass is lumped with the piston assembly to get the total reciprocating mass.

Another request - would be nice to know the X-section area of the production con rod at the C-to-C distance 2/3 up from the big end. This seems to be the approximate location where most connecting rods fail from fatigue.

Knowing the X-section area and alternating force applied, we now know the alternating stress applied to the con rod and can check this against the material properties to see if we are encroaching on the fatigue limit.

Let's use the 3.2 L engine as the specimen. I've wanted to do this and now have a good reason to do so, but didn't want to disassemble my engine just for the measurements. As an engineering consultant in the engine business, I have the software to make quick work of this problem and with the help of Chris on the metalurgical end, I think a real answer can be had. Should be fun!

Dave,

To model it well you may need to estimate the twist that occurs and factor in this value. Also peak cylinder pressures may help.

I would guess a peak accel in the order of 5000g as maybe somewhere to start.

By developing a loading spectrum it may be possible to do some rainflow analysis and come up with a life estimate.

We used to run nCode Software on our Fatigue Testing machines but this was during the time when it was being developed at British Rail in the late Eighties and early Nineties.

Compressive loads are not as bad as tensile loads (for example on high rpm deceleration with engine braking) where mass of rod/piston assembly counts the most.

Hence could do the calculation even without knowing the peak cylinder pressures.

I am not sure what you mean by this statement but it needs some examination.

If you are considering crack growth then the effect of crack tip closure on da/dN needs to be considered and by testing in a tension/tension test could result in significant crack growth rate errors.

Crack initiation is far more complex and is generally driven by resolved shear stresses, which when the become critical cause the intrusion and extrusion of slip bands that lead to the development of a crack. Twist and Euler Buckling contribute to this behaviour and are both caused by compressive loading.

If we consider a crystal that has a favourable orientation then tensile loading causes a positive resolved shear stress and when this reaches a critical value slip will occur.

The action of this slip will be to produce a residual stress and due to the continuity of the stress field within the crystal the residual stress relieves the positive shear stress not only locally but across the crystal. This means that the positive shear stress is very localised during tensile loading.

Effectively reducing relieving a positive shear stress has a similar effect to increasing the negative shear stress and during compressive loading the greatest negative shear stress will also cause slip.

This slip then relieves the negative shear stress which effectively strengthens the positive shear stress and makes slip in the tensile direction much more likely.

As this process repeats the magnitude of positive and negative slip increase monotonically with increasing cycles of loading and produce a slip band extrusion.

This is the beginning of crack initiation.

I do believe that the academic world has long accepted that fully reversed loading is more damaging that a tension-tension case particularly in terms of initiation.

I would agree that looking at models with and without compressive loading would be interesting but I would think compressive loading will affect fatigue life.

Developing a realistic loading spectrum would be a good first step

http://rs1234.wz.cz/img_m/ART_166.pdf

Hi,
I miss only the weight of segments.
The set was weighed with a precise balance of laboratory in 0.01g.

Piston : 443.95 g 98mm Mahle
Rod: 508.69 g R/R of LN engineering. Origin rod : 660 g
Axe: 127.94 g
Clips: 0.91 g/u (2)

Total: 1082.4 g

;)

I think Raceboy is saying, no one here makes enough power to bend a rod.

RPM? Sure, someone can mechanically overrev.

I "money shifted" mine hitting somewhere in the 8k range with stock rodbolts, and no issue with throwing the rod(s) out though....

Con Rods do bend - I think there can be no doubt about this - the forces generated may not cause plastic deformation but this debate really centres of elastic behaviour and my point is that tension, compression and bending all need to be modelled if we are to obtain good results and be able to come to good conclusions.

The entire debate about the 'life' of a Titanium Rod is to due to the accumulation of damage at a microscopic level.

If you ignore reversed loading during compression you will almost certainly overestimate fatigue life as you will lose around 50% of the stress amplitude, bending is part of this issue as is buckling and twisting which depends on the design and tends to by why H Beams are favoured in Turbo engines.


The stresses in rods generally has a basic form which will change depending on several variables but the basic shape of the distribution doesn't change.

http://i197.photobucket.com/albums/a...psnwcndefs.jpg

I have been interested in the design of con rods for some time and we have manufactured our own con rods for 2.0 litre race engines for the last 3 years and would like to make a wider range of products within the next 12 months.

If normal combustion is bending rods, isn't detonation a lot harder on them?

Love to hear 8K, any video clips?SmileWavy

I would think so but bending stresses should still be within the elastic limit - the closer to the elastic limit the more likely it is that fatigue damage will accumulate and the shorter the life.

The models being used by the leading research companies could add detonation to their load spectrum and estimate the impact of this problem.

It is still fair to say that fatigue failure of 911 rods is very rare so the design is conservative but the Ti rods that started this debate are more of a concern.

If the Ti is manufactured under controlled conditions and I would assume that material purchased to AMS specifications should be consistent then it is likely that the quoted 'life' is conservative.

If the rod is then not used to the limits that were used during the design stage it may be capable of much greater than the quoted life.

Without knowing all the details it is just difficult to judge.

It was a "money shift", mechanical overrev.

Redlined 3rd, then decided I'd like to put it in 2nd instead of 4th.

Damn G50's shift so well! ;)

Starting to wonder if motors that have thrown a rod (any motor, not just 911's) in the past possibly were from detonation?

Of course you have your standard oiling issues and overrevs.

Personally, I've never seen an actual rod failure from detonation in any engine (going back to 1962 when I started). :)

Usually, pistons/ring fail before stresses build on rods prior to outright failure.

JMHO, but the majority of rod failures stem from lubrication/bearing issues which trigger seizure and I've seen my fair share.

so you're saying to fix any stretching from over-rev, just turn up the boost?

http://forums.pelicanparts.com/suppo...eys/icon26.gif

That's what I've always believed too...

Pat - Thank you for the component mass information. I’ll start assembling the input data based on these numbers. I noticed that the piston mass is for a 98mm bore – the 3.2 L engine is a 95mm but what you provided will be fine for my purposes. I’ll use the original rod mass of 660g and 1/3 of this (220g) is considered the reciprocating mass contribution of the rod. Added to the piston assembly, this totals 793g for the recip mass. I have all the rest of the geometry of the 3.2 engine at hand, so just now a matter of preparing the input file for the analysis.

For cylinder gas pressures, I use the Vibe heat release model which has proven accurate to within a few percent of not only predicting the peak cylinder pressures encountered during normal combustion, but the pressure at every crankangle increment during the power stroke due to a realistic heat release calculation. Inputs are delivery ratio (volumetric efficiency) compression ratio, A/F ratio, engine speed, and target BMEP. Cylinder gas pressure force applied to the piston dome area is superimposed over inertia loading, and this comprises the basis for the dynamic model. Most of this input info I’ve collected on the 911 engine over the years or know where to go to get it.

Chris – When you say twist that occurs, do you mean a torsional deflection of the con rod? I realize this can and does occur in actual running engines but my code does not take this into account. Also, the radial outward force in the plane of the con rod that causes bending is not accounted for as this is usually small in magnitude as compared to the recip forces, and is ignored. However, I can factor this in as a separate calculation and it may become significant with engine speeds of 8000 rpm.

I’ll work on this as a “hobby project” over the next few days and will present results ASAP. Can’t drive or work on my 911 at the present time (winter in Wisconsin!) so I’m deskbound.

Still working on it - should have something to post in 24 hours.

Dave,

This may be interesting

http://www.eng.utoledo.edu/mime/faculty_staff/faculty/afatemi/papers/2006JMESShenoyFatemiVol220PartCpp615-624.pdf

Chris - thank you for the technical paper link - I will peruse this over the weekend - looks excellent. The analysis that follows does not consider the bending forces in the plane of motion of the connecting rod. Will have to provide this as a separate calculation.

This thread sparked an interest in learning more about the inner-workings of the Porsche air-cooled engine – namely, what kind of loads are the connecting rods subjected to in a modified engine. Looking at a worst-case scenario would allow one to put some numerical values on the maximum tensile and compressive loads the con rod feels. Undoubtedly the Porsche engineering department went through a detailed analysis during the design phase of the connecting rod and to my knowledge, the factory rods prove very durable and failure-free, even in high-revving, competition engines. Nevertheless, it’s just interesting to be able to assign numbers to what is generally accepted to be the most highly stressed mechanical component in the engine.

I selected the 3.2 liter engine for this study. Here are the following specifications used for this analysis:

Bore – 95.0 mm
Stroke – 74.4 mm
Rod Length – 127.0 mm
Comp. Ratio – 9.5:1
Ign. Advance – 26 deg btdc
Mass Piston Assy – 572 g
Mass Con Rod Small End – 220 g
Mass Con Rod Big End – 440 g

The numerical model consists of a dynamic simulation of the slider-crank mechanism (crankshaft-conrod-piston) utilized in virtually all piston engines. Knowing the mass of the various components and the crankshaft rotational speed, it is straightforward to predict the inertia-induced forces in any of the components of interest. Superimposed on the inertia force is the cylinder gas pressure loading that is applied to the piston dome. Having accurate gas pressure data during the compression and power strokes, is paramount in obtaining a realistic dynamic loading of the components during the 720 degree rotational interval of the 4-stroke engine cycle. Such a simulation is standard practice for the design of a new engine, yielding important information on the dynamic loads of both the wristpin and crankpin bearings.

Not having access to any actual cylinder pressure data taken on the 3.2 engine under laboratory testing, the techniques given in Ref. 1 (Vibe combustion model) are employed to generate a theoretical gas pressure file. Inputs, in addition to engine geometry, are 1) compression ratio, 2) delivery ratio, 3) ignition timing, 4) air/fuel ratio, 5) ignition delay, 6) burn rate in crankangle degrees, 7) engine rpm and 8) BMEP. 1, 3 and 4 are known from Refs. 2 and 3. 2, 5 and 6 are based on experience and collected data on other air-cooled, 2-valve, naturally aspirated engines. Engine rpm and BMEP are based on the target performance we arbitrarily select for this study. Based on an earlier study posted several years ago (in which I probably confused more than informed, sorry), I looked at realistic BMEP outputs of the 911 engine in various displacements and states of tune.

http://forums.pelicanparts.com/911-engine-rebuilding-forum/645505-accurate-approximation-porsche-911-engine-power-output.html

For peak power I chose 188 psi BMEP @ 6100 rpm where the 3.2 L would be producing 280 bhp. For the peak engine speed desired of 8000 rpm, we’re way past peak power, so an educated guess has to be made regarding output here – I think 200 bhp is realistic with a corresponding BMEP of 102 psi.

Running the combustion simulation for the peak power point results in the following PV diagram with an air standard cycle shown as reference to a “theoretical maximum”:

http://forums.pelicanparts.com/uploa...1489783084.jpg

The sharp discontinuity at max volume is due to the “instantaneous” blow-down of the cylinder pressure, as the exact opening timing of the exhaust valve and associated pressure history is not considered nor is necessary for this analysis. Highlights here are maximum cylinder pressure developed – 923 psi.

With cylinder pressure now approximated, this can be applied to the kinematic model to obtain the force applied to the connecting rod beam, this being a function of the inertia force of the recip assembly with gas pressure force superimposed, for a full 720 degrees of rotation (1 full engine cycle).

http://forums.pelicanparts.com/uploa...1489783172.jpg

0 degrees of crank angle corresponds to tdc at the beginning of the inlet stroke – here the maximum tensile load is encountered as there’s no appreciable gas pressure on the piston dome to offset the inertia load of the recip mass. 180 degrees is bdc of the inlet stroke – moderate compressive loading due to inertia. 360 degrees is tdc on the power stroke – notice that prior to tdc the tensile force rapidly changes to compressive loading due to combustion initiating and providing the dominating force around 20 degrees atdc. 540 degrees is the exhaust open point where the cylinder “instantaneously” blows down to atmospheric pressure with very little change in loading. 720 degrees is the end of the exhaust stroke, with tensile load again at a maximum.

Here is the important take-away on the con rod forces at the max power point:

Force Max Compressive – 6790 lbf
Force Max Tensile – 3492 lbf
Load Cyclic Frequency – 50.8 Hz (@ 6100 rpm)

I don’t know the rod X-section area so can’t calculate the stress. If someone can provide an accurate approximation of the beam area at the 2/3 center-to-center distance (up from the big end) then we know the cyclic stress.

More to follow, including references.