conrod stress during each stroke and failure 2

[tomaschan]

Hi, as far as I know conrod is stressed by tensile, compression and bending stress.
I would be interested what stress is imposed on conrod at each particular stroke of 4-stroke engine and when is it most critical.
And at which place of shank (closer to big end or small end) is higher probability for conrod to break and why.

 

[Tmoose]

highest tensile force is at TDC of the exhaust stroke. At high rpm it can be the highest loading by far.
The Geometry of the con rod determine the stress under a particular load. That includes the basic form plus every detail of how the finished rod was manufactured, also including if there are any material defects.

Small block Chevies of the 70 often showed "indications" in the notch where the rod bolt head seats.

http://remanufactured.com/images/350ChevyReconditionedRod.jpg

Sometimes but much less often "indications" would appear somewhere along the I-beam, usually starting from a notch or deep scratch left over from original manufacturing, or a notch filed into the I-beam as a means to "number" the rod during a previous rebuild by a badly informed mechanic.

Studebaker issued a factory bulletin about 6 cylinder rods that were breaking in service. The break originated in the rod bolt head notch.

Harley 74 con rods often showed indications in 2 or 3 particular regions of the opening machine in the forked rod. Sometimes but much less often "indications" would appear somewhere along the I-beam, usually starting from a notch or deep scratch left over from original manufacturing.

Rods modified "for racing" sometimes end up with severe stress concentrations and may end up "weaker" than well prepared "sotck" rods.

http://www.linkcycles.com/ebaypics/IMG_1087.jpg

The "indications" can range from barely detectable with wet method mag particle inspection, to so deep as to be visible with the naked eye, to one end of an ugly shard among the battered remnants of a rod that finally failed catastrophically due to fatigue.

 

[ivymike]

you said four stroke but you didn't mention whether it was turbocharged. rod tensile loading is negligible on some turbocharged 4-stroke engines.

 

[tomaschan]

Thanks for remark, interesting rods for racing...
Regarding stress, yes, that was my assumption - highest tensile stress at TDC of exhaust stroke, but isn't compression stress at TDC of compression stroke higher due to high pressure on piston? What about stress at BDC? And why should be there negligible tensile stress in turbocharged engine?

 

[ivymike]

regarding the turbocharged one, because there is always pressure pushing the piston down.

In a gasoline engine at high rpm, tension at TDC-exhaust is a high number. cylinder pressure is not typically very high on a gasoline engine, so lower magnitude stress due to firing at TDC-compression when compared to a turbodiesel which may have double the peak cylinder pressure. The acceleration-related force which causes tension at TDC exhaust also exists at TDC-firing and counteracts the firing load.

 

[Tmoose]

4 stroke TDC exhaust.
If it was possible for boost to pressurize the cylinder well before TDC (when the intake has not yet opened) and through, during and after TDC when the ex valve is open as well, 15 psi would exert less than 200 lbs force on a 4 inch diameter (12.6 in^2) piston.

According to this online calculator (

http://2.3liter.com/Calc2.htm#PistSpeed

) a stock 1976 350 Chevy (3.48 " stroke, 5.7" long rod) piston at 5000 rpm experiences about 1600 gs.

If a stock piston with pin and rings weighs 650 grams (1.4 lbs) or so then at 5000 rpm the rod has to exert 1600 x 1.4 = 2300 lbs to haul the piston to a stop and make it reverse direction. That force estimate is at the wrist pin centerline. The crank journal/rod big end bearing also has to haul the rest of the con rod back down the hole too, and some portions of the con rod are the load path for that feat.

2300 - 200 = 2100 lbs tension taking full credit for boost pressure.
It would take 185 psi (~ 12 bar) cylinder pressure at TDC to balance the inertia load. That is in the ball park for TDC pressure on the firing stroke, before full cylinder pressure is developed at 10-20 degrees ATDC or so.

i don't have any info for heavy diesel components at 1800 rpm.

 

[ivymike]

If it was possible for boost to pressurize the cylinder well before TDC

Don't forget that exhaust pressure can be elevated on a turbocharged engine.
Take a turbodiesel operating at 900rpm (rated), with 280mm bore diameter and a recip mass around 60kg, it only takes 1.52 bar to counteract all of the acceleration. At full speed and 0% engine load you have about 0.95 bar at the end of the exhaust stroke, so not completely balanced (you're left with about 3500N to worry about). For comparison, the rod compressive force due to firing at full load (after you subtract inertia, as if it matters) is 1,130,000 N.

There are some large-ish turbocharged 2-stroke engines wherein the connecting rod large end only supports the bottom of the piston pin - it doesn't go around it - the piston doesn't fly away because there is always enough cylinder pressure to hold it down.

 

[140Airpower]

ivymike, that does not sound like an auto engine. 60kg of reciprocating mass is huge. The problem is that g forces increase with the square of rpm, being 4 times as great at 1,800 rpm as at 900 rpm and 64 times as great at 7,200 rpm. For auto engines, especially normally aspirated racing engines g forces absolutely dominate the stress profile.
The two areas of greatest strain due to tensile stress in a connecting rod are the neck, just below the piston pin in the typical rod and the big end hoop. Failures due to overload, where there are no particular stress risers, usually occur at the neck where the rod simply snaps from being pulled apart and at the big end where the stretching of the pin eye, along the long axis of the rod, into an oval shape pulls the sides in where they pinch the crank throw, causing high pressure, oil film failure, high friction or seizure of the crank throw, rotation of the big end with the crank and consequent bending of the rod, breakage at the neck... followed by generalized mayhem.

For low rpm engines and for highly supercharged engines or for engines experiencing detonation or hydraulic lock, compressive forces usually dominate. Rod failure is then usually due to column collapse or bending. Buckling forces are 4 times as great in the plane of pin rotation as in the plane of the pin axis (see Euler Column Theory). So, the rod is generally wide in the plane of rotation and narrow in the axial plane. But, collapse in the plane of rotation is preferred as the direction that will cause the least engine damage. Consequently, rods are usually designed to be somewhat less than 4 times stronger in the plane of rotation.

For very high performance rods where strength to weight is more critical, design details are also more critical. This is where you really get into the questions of fatigue life and reliability; bolt notches compared to threaded bolt holes, machined and even polished surfaces vs as-forged surfaces, rod cap pinning or serrated mating surfaces, overall shape as well as contour details with respect to stress distribution, high strength vs light weight materials, cast vs forged vs billet, etc.

 

[ivymike]

that does not sound like an auto engine
I didn't see anyone say we were talking about car engines?

 

[dicer]

If you want to see what kind of rod survives at high rpms look at high reving motorcycle engines.

 

[tomaschan]

Yes, I forgot to mention I was thinking of some "better than ordinary" car engine, let's say 4cyl 2liter gasoline, where turbo boost of 1-2bar wouldn't help a lot in decreasing of the tensile stress caused by huge acceleration of reciprocating mass at high rpm, I think.
However, I would be interested to see the progress of compressing and stretching forces (imposed on conrod), when do they swap during whole cycle...

 

[patprimmer]

Well from some previous posts it can be seen that it varies greatly with piston weight and piston acceleration rate

Regards
Pat
See FAQ731-376 for tips on use of eng-tips by professional engineers &

http://eng-tips.com/market.cfm

for site rules

 

[ivymike]

(correction to my post above "connecting rod large end" was meant to say " connecting rod SMALL end" - not that it matters)

so if we're talking about high speed gasoline engines, forget the turbo comments, the closed throttle max speed condition will be worse. (I think that was mentioned already)

This ought to be informative for you:

https://www.forging.org/system/files/field_document/DynamicLoadAnalysis.pdf

 

[140Airpower]

Ivymike, the category is automotive and the OP was asking about car engines. That's why I pointed out that what you were describing was not automotive. But, what you said was perfectly true of very big diesels and was useful to illustrate that the balance of stress depends on the rpm range.

This also is informative:

http://www.autosteel.org/~/media/Fi...ility and Optimization of Connecting Rods.pdf

In a previous discussion on rods someone posted a video showing FEA analysis of a rod going through the complete 4-stroke cycle. Does anybody have that reference?

 

[140Airpower]

Here are some Youtube videos of dynamic stress analysis of con rods. In the first you can see floppiness of the rod which is greatly exaggerated, but it illustrates that bending results from inertial forces on the mass of the rod itself. The big end oscillates like a watch balance wheel while the shank whips back and forward. These forces are minor compared to the stretch and compression, but they add to those and perhaps can help initiate a collapse under compression.

http://www.youtube.com/watch?feature=endscreen&v=uTpRfREYa9Q&NR=1

http://www.youtube.com/watch?v=RMLXnLqvvI8

 

[Lou Scannon]

Ivymike, the category is automotive and the OP was asking about car engines.

140, I missed where the category is narrowed to automotive; and the OP didn't clarify "car engine" until after Mike's informative post about boosted 2-stroke engines. Your subsequent post was very informative also, BTW.

"Schiefgehen will, was schiefgehen kann" - das Murphygesetz

 

[Tmoose]

I really expect designers have enough to worry about keeping the rod together under prolonged service at max power. Catastrophies initiated by rod bearing failure, unless the bearing failure resulted from Big end distortion, are the worry of the lube system group. A 4 stroke rod will accumulate 10,000,000 exhaust cycles in 50 hours at 6000 rpm. For a rod to break in operation is almost certainly a fatigue event, and that requires cyclic tensile stress. Compressive forces just won't do it.

 

[ivymike]

"Automotive" includes ships, planes, trains, etc...

 

[Lou Scannon]

Compressive forces just won't do it

agreed, when the engine is operating within design limits. I've seen rods that have been abused via knock/overboost that have yielded compressively.

"Schiefgehen will, was schiefgehen kann" - das Murphygesetz

 

[ivymike]

For a rod to break in operation is almost certainly a fatigue event, and that requires cyclic tensile stress. Compressive forces just won't do it.

In the end it is not that hard to design a rod to survive either tensile or compressive loading unless you're trying to get very close to the edge and using a non-ferrous material. It has in the past been done very well without the aid of computers - doing hand calculations of stress at a few key areas was/is enough. The funny thing is - you'll have rods that break anyway, once in a long long while ... and that's when the job gets interesting.

Assuming that there was not a failure of another engine system to precipitate the rod failure (like overfueling leading to piston seizure, or a bent cooling jet, etc), manufacturing defects (forging laps in particular) and assembly errors (mostly to do with bolt tightening) have accounted for most of the rod failures I've had to sort out. Rod BOLT failures, on the other hand, can be extremely tricky.