Connecting rod stress 1

[dodge8564]

Can someone please explain how load is transmitted through the drivetrain to the connecting rods?
I am in a discussion in another forum about how an engine needs stronger connecting rods for towing, even if the engine never reaches above 3500 rpm.
Another person is arguing that load is not transmitted through the drivetrain to the engine itself, and the only reason to have stronger rods and bearings is for high rpm's.

 

[enginesrus]

"Can someone please explain how load is transmitted through the drivetrain to the connecting rods? "
Yeah guess we did miss the question.
The load starts at the tire contact radius, and is transmitted through the axle, to the differential gear set, to a drive line or propeller shaft if it is hotchkiss drive, then to the transmission then to the crankshaft and of course then to the rods.
All gear sets in between are torque mulitpliers. And in the end, in theory, you see close to the same HP at the rear wheels that the engine puts out. Actually though it isn't because of losses.

 

[Rob45]

The greatest load on a connecting rod is inertial load, occurs at top dead center on the exhaust stroke, and produces a tensile stress, not a compressive stress.
This is true regardless of engine load, and depends ONLY on the RPM for a given engine.

 

[ivymike]

The greatest load on a connecting rod is inertial load, occurs at top dead center on the exhaust stroke, and produces a tensile stress, not a compressive stress

Is it fair to say, then, that you believe there is no such thing as an engine where the firing load exceeds the inertia load on the conrod?

 

[patprimmer]

I think that is approximately true for normally aspirated engines, but once you get into high boost and/or methanol and nitromethane fuels, the compressive loads can get very high, but this has nothing to do with the original question.

The highest load on the bolts is at maximum RPM with no load. ie a closed throttle as the vacuume in the inlet manifold helps to stretch the bolts by reducung the pressure on the piston that might otherwise help it to change direction at TDC. I think the highest stress is at the highest piston acceleration rate on its way down the induction stroke. That position changes a bit depending on rod to stroke ratio, and I don't see a need to calculate it accurately, but it is definitely well after TDC.

At TDC, the piston is instatainiously inertia free, as it is instantainiously stationary.


Regards
pat

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[ivymike]

That position changes a bit depending on rod to stroke ratio, and I don't see a need to calculate it accurately, it is definitely well after TDC.

Hey Pat, it might be worth double-checking your logic on that one. The piston is instantaneously motionless at TDC (excluding secondary motion), but it is certainly not without acceleration.

 

[patprimmer]

Yes

A moment of brain fade, it must be subject to acceleration to stop and start again.

It's to long since I studied simple harmonic motion and I must admit I forget how to do the sums.

I know from experience that rod bolts most often break as you lift of the throttle at maximum RPM.

The lift of adds manifold vacuume to the forces on the piston, so the maximum force must be after the exhaust valve closes, unless the pulse tuneing and gass inertia pulls a stronger vacuume in the exhaust than the closed throttle does in the inlet. I can't imagine this ever being so, so maximum suction on the piston must be after the exhaust valve closes, and the resultant force from vacuume over the piston and inertia must therefore be greatest after the exhaust valve closes, otherwise you would not get more failures on a closed throttle

Regards
pat

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[enginesrus]

Pat, just curious. On the bolt breakage. What engine and what rpms? How many bolts broke?

 

[crashbox455]

i don't understand one thing- if an unloaded vehicle under WOT reaches the same cylinder pressure as a loaded one under part throttle, why does the loaded one detonate?

resisting the motion of the piston will increase cylinder pressure, perhaps beyond the "shock" tolerance of the rod? i'm not a structural engineer, so maybe someone who is can explain why i'm wrong. but it seems to me that putting the vehicle under load will spread the peak pressure out over an extra degree or two of crank motion, possibly causing premature rod failure.

anyone care to counterpoint?

 

[patprimmer]

A long time ago, (197?) once only, the rest are only anecdotal

Roadracing 350 Chev LT1

7500 rpm stock (pink spot) rods, 3/8" ARP bolts, endurance races.

The engine peaked out at 7000 rpm, but the driver had a habit of changing down comeing into a downhill corner.

Only one bolt broke that I saw, but a number of comments were made by others who saw it, that it always happens as you back of and it goes lean. I knew the goes lean was wrong and has nothing to do with it, but the guys who commented were much more experienced than I, although not nearly as scientific.

I always attributed it to the extra vacuume as the throttle closed at very high rpm.

Since then I always used aftermarket rods and 7/16" bolts. No more failurs.

Regards
pat

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[patprimmer]

The engine at WOT is always under load, even if it is it's own inertia. Low external load will see a very rapid increase in rpm as the cylinder pressure overcomes the inertia load. If there was no load, rpm increase would be instantainous instead of very rapid.

The initial question was not about no load vs loaded to stall, it was about the variables from the load of an unladen car to a heavily laden car, say an increase of 1400Kg to 3000Kg as an example.

I have never ever seen a standard rod in a standard engine built since the 1960's fail in compression, unless hydraulic locked or seriously flawed in the first place (actually I have never seen that either, but consider it possible).

I have seen engines that have been hydraulic locked. They pull every head studor blew the porcelain out of spark plugs, but still not bend rods.

I have seen rod bolts break, once in an engine I was involved in (see above), and in other engines, in which I was not involved.

I have never seen an engine failure that could be directly and solely attributed to a rod or rod bolt failure in a standard engine in an everyday drive car. Any failed rods in these cases had other damage such as spun bearings, broken gudgeons etc

Regards
pat

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[GregLocock]

My partner in crime once assembled a diesel engine from two donors. He had not realised that the block height and the conrod length were matched, and used the block from one and and the rods from the other.

On reassembly the engine ran OK, but very quickly blew the head off, by stripping the head bolts.

We bodged it back together only for the same thing to happen again.

This time he looked around carefully and noticed that the serial number stamped into the piston's crown was clearly visible on the head (or vice versa). The piston had been hitting the head hard enough to break the head bolts.

The con rods were fine.





Cheers

Greg Locock

 

[ivymike]

couple of questions -

* Is it necessarily the case that impacts between the piston and the head cause the conrod to be loaded in compression (which is what I think you were implying above)?

* Are you guys (pat and greg) trying to tell us that any conrod that is capable of withstanding the tensile loading in a (naturally aspirated, if you must) engine can withstand the compressive loading that it experiences in the same engine? (presumably the reason you would assert such a thing is to counter the previous statement by unterhausen?)

 

[patprimmer]

If a piston is hitting the head hard enough to pull the studs, I think it is safe to presume the reaction is transmitted back through the piston through the pin to the rod.

I know that whenever I walk into a closed door, my nose certainly compresses, even if I do not hit it hard enough to pull the hinges off.

I certainly am NOT saying that because the tensile is adequate, this means the compressive is also. This would imply that strong rope could work as a con rod, as it has good tensile, shame about the compressive though.

I am saying that in all engines that I have seen, used or have relevant information on, which is quite a few, the conrods are built with such a safety factor in both tensile and compression, that they never ever fail, unless something else happens to break them.

For tensile, they simply need enough cross sectional area of a suitable strength steel. Design (appart from notches) does not matter.

For compressive, they need a beam with depth of section in height and width (if you presume that length is pin to pin direction) so as to resist bending.

I see no real reason to counter much of unterhausen's post. despite him saying the question was not answered, he gives much the same answer that is repeated over and over again, with slightly different symantics.

The original question is well and truely answered, and this thread is starting to get really hyperthetical and boreing, so I don't think I will have anymore to say on the matter

Regards
pat

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[Rob45]

To InHiding, who wrote:
"Is it fair to say, then, that you believe there is no such thing as an engine where the firing load exceeds the inertia load on the conrod?"

The short answer is, 'Yes.'
Note all the other stories above about rods failing when the throttle is closed and the cylinder pressure drops dramatically.

And 'beliefs' have nothing to do with it; I've done the math, and instrumented and tested the engines.

 

[Tmoose]

I've heard several stories about rods failing at closed throttle.

There are a few pieces of those stories I have never managed to fit together well enough for my satisfaction.
1 - fatigue failures (cracks) can take hundreds or thousands of heavy stress cycles to progress from initiation to the big bang grand finale. With the aid of the magnaflux process I have seen dozens, if not hundreds of rods and crankshafts with cracks in the expected high stress areas. Those cracks have varied from so tiny as to be removed by a little polishing, to so deep and wide that it was only half a joke to say that the owner had turned off the ignition on the engine revolution just before the part broke in half.
2 - I've seen a few busted rods and crankshafts and valve springs. The crack Always (even the battered parts) showed the text books signs of crack progression. Really, the only question might be how many cycles it took for crack to hurry thru the last part of it's final journey.
3 - I figure The tensile force on a 4-stroke connecting rod at the exhaust TDC is always un-opposed by significant cylinder pressure.

So, I have long felt that heavy reversing stress cycles accumulate quickly at full throttle max rpm. They may accumulate at most twice as fast at closed throttle max rpm.
So, in my mind, 2 minutes of max rpm full throttle is much like 4 minutes of max rpm closed throttle. Since the forces vary as speed^2, an extra 500 rpm (on a bad down shift) is probably a much greater danger than simply 'backing off."

Well prepared parts (magnafluxed, well radiused and shot-peened) have a nice fat bank account of useful stress cycles to spend before failure. A this point it would surprise me if the "fatigue checks" only get cashed when the throttle is closed. If I am missing the reason part throttle stresses are SO HIGH, then it would seem likely that several rods would often break at once, and maybe all the rods would break at the same time once in a while.

 

[Skeebo]

Pat finaly some one said it.The highest cyl press. is when the throttle is snapped shut after WOT.This causes a Hydrolock affect in the cyl.You can prove this by watching the KV scale on any secondary ignition scope.
This is why you cant run alluminum rods on the street they will fly apart on decell.as far as having to spend big bucks on a set of rods just debur all edges and sides verticaly and pollish the rod will eliminate any stess fractures from starting.we ran stock rodsfor ten years in the national tractor pullers asso.never broke a rod.we ran twin 396 chevy's approx 1200 hp each.the only break downs where usualy caused by valve failure.
thanks
Todd Sternberg

 

[GregLocock]

I find it hard to believe that. Where does this pressure come from?

Look at the load on a piston as it approaches TDC.

Acceleration at TDC is roughly r*w^2, or at 6000 rpm with 100mm stroke, that is 0.05*(6000/60*2*pi)^2

So that is a force of about 11 kN on a 550 g piston. In the absence of any gas forces that is all supplied, in tension, by the rod.

If it is compressing gas then there is downward pressure of CR*piston area*atmospheric pressure, or say 9.5*pi*.046^2*101325= 6kN, on the piston. Therefore the load on the connecting rod has HALVED at full throttle compared with the unloaded case at the same speed.

So when the old guys talk about the intake charge cushioning the piston they were right.




Cheers

Greg Locock

 

[ivymike]

I've done the math, and instrumented and tested the engines.

That's a bold statement - to prove the assertion that no such engine exists, you'd have to test every engine in existence, wouldn't you? Blanket statements are dangerous; I didn't figure I'd have any "takers" on that one.

See if these figures work for you:

peak cylinder pressure 170bar
rated speed 1800rpm
cylinder bore 155mm
stroke 165mm
rod length 310mm
piston and pin assy mass 7.21kg

conrod load due to pcp: ~320kN
conrod load due to inertia at tdc exh: ~27kN

I know which one I'd be more worried about. There may be at least one more such design out there.

My point is simply that one must check both buckling and tensile loads when doing classical checks on a connecting rod. Sometimes one is big, sometimes the other. It's often easy enough to guess which will be the big one, but if you've done the calculations once, you can run through them all in about 5 minutes with a hand calculator, so why not do ALL the checks instead of just waving hands and pointing thumbs?

 

[ivymike]

...for accn at TDC, I prefer 0.5*(stroke)*(w^2)*(1 + 0.5*stroke/rod_length). With common ratios of rod/stroke, the number is about 25% higher than the one you'll get with the abbreviated formula above.

 

[framorgal]

I think that many of the participants on this thread are confused whith the concep of "towing". Really is the same if the engine is used for towing or if is instaled into a test bench performing a Power curve. If the engine runs stationary at a given speed -WOT-, this means that there is an equilibrium between the power developed for the engine and the power consumed by...wheels, propeller, dynamometer... is the same!, if the engine doesn?t runs stationary and decelerates, this means that is needed more power from the engine, and if is possible the system (engine-consumer) stabilize at a different engine speed, if not the engine stalls.
About the rod mechanical loads, you must to consider a complete engine cicle -720?- and apply to the rods THE SUM of all the forces calculated each degree (for example):
Piston inertia, Gas forces, Rod inertia, bearings resistence...
In the worst situations:
Max. engine speed (with and without load), max. torque speed (full load), sudden deceleration at max engine speed, sudden acceleration whitout load
All that forces produces traction, compresion and bedding in the rod, and after you must consider that the real system is not static (crankshaft torsional vibrations among others) and the max. peak pressure can be higher (Knock), for this reason you must to apply a safety factor if you want to run sure.
In sport you usually play with this safety factor (taking mass fron the rod) when you need an high rev. engine, or when you want sudden acceleration, but if you use the engine in almost stationary speed (like towing) the rod mass doesn?t represent any problem, dont spend your money in lightweight rods.

Sorry for my English, i'm learning