[Lee]

I started using a stretch gauge about 25 years ago. Some of the bolts can be REALLY tricky to get to, with only about 1/4-turn possible without repositioning the wrench. And unless you are a gorilla, you are NOT going to get the proper stretch with a normal boxed-end wrench! I've got some extra-long Snap-On box-end wrenches I use, and still have to wrap rags around the end to keep from making my hands sore.

I've refined my technique over the years. I now measure a rod bolt, remove the gauge and torque it to (for example) 50 - then remeasure the rod length. Then add 5 and repeat until I get the desired stretch. Once I've determined what the magic torque number is, I repeat for another rod bolt - but I again measure the rod before and after to verify. I check every bolt that way, and tweak the magic number as needed as I go along.

[Gach]

There’s allot of good videos on this whole subject from professional engine builders. I particularly like this one seems to be the simplest way using a good stretch gauge

https://youtu.be/sxW7TJTWoOY

[Gach]

Another one using torque angle torque wrench

https://youtu.be/im1YgvBvV_8

And another one using regular all torque wench and stretch gauge.

https://youtu.be/kA1elSAnE7g

[Gach]

Are all your rod bolts same length have you checked them.

https://youtu.be/0CgqeJ-V094

[Shiny]

Originally Posted by Elarson

Great questions. But you're flirting with getting deep into bolted joint design which could easily whirl out of control. But here's the overall rules:

Thanks for taking the time to answer. I think you avoided getting too technical very well.

1) the rule of thumb for a typical fastener is to tighten it to no more than 75% of yield. Note that if you use a torque wrench that is +/- 25%, you will come close to yield (about 94% of yield) but not exceed it. For example, a Grade 8 bolt is 150,000 psi ultimate, approx. 132,000 psi yield so the target preload would be 132,000 x .75 = 99,000 psi.

Got it, thank you. Good graphic about the effect of variation. I guess I never thought friction could add that much variability in clamping force. Wow...

2) If the bolted joint does not separate, the bolt load will not go above preload. Think about bolt stretch vs. load; and the only way to stretch a bolt farther than it's installation stress is to open up the joint. For example, if the bolt is preloaded to .005" stretch and operating load makes the joint bounce open by .001", then the bolt stretch is now .006 and the bolt load is 20% higher than the preload. If it's in something like a connecting rod that has a cyclic load, then the extra 20% stretch (load) becomes an alternating stress and that's what kills a fastener in fatigue. So a properly designed bolted joint never separates and the bolt never sees higher stretch (load) than it's installation preload. Which leads to the next rule of thumb:

This is not obvious to me. I guess I expected more of a "free body diagram" exercise where the bolts actually compressed the rods to some degree. In other words, the rods and caps are loaded and compress to counter the bolt expanding. What I'm struggling with is why the stress acting on the rod (ie inertial loading from piston and rod mass) doesn't add stress to the bolt? I get that the inertial loading would have to exceed the clamping force to cause complete separation of the cap, but wouldn't inertial force cause the bolt to elongate since it adds to the clamping stress? I picture the bolt elongating and the compression force within the rod and cap being reduced, so they also stretch.

3) If practical, it's always safe to design the bolt preload to be double the anticipated operating load. Then the joint never comes close to separating. So in the previous Grade 8 example, the operating load should be equivalent to no more than 99,000/2 = 49,500 psi. Aircraft engines and nitro funny car engines don't typically have that luxury and the joint/fastener design becomes more critical.

Hopes this helps without getting too deep in the weeds.

Eric

Yes, that is good info and makes sense from the standpoint of avoiding separation.

But I'm still thinking inertial loading will cause cyclic stress on the bolt, even if it is clamping and the joint design prevents separation of the cap and rod. I'll have to think about this more but will try not to whirl totally out of control!

I have used Miner's rule in the past to effectively "sum" cumulative fatigue damage from complex loading like vibration stress, but I need to think more about the "smaller cyclic stress" on top of a preload..

And all this makes me wonder why a stock rod bolt would ever fail. Is it all about the rpm and resulting inertial loads overpowering the clamping force?

All interesting so thanks again.

Mike

[djustice]

Was just watching a video about this very thing from a former writer on high performance pontiac magazine. he is testing rodbolt torque vs strech and checking the results with a micrometer . To see if they need to be resized (on a ford rod) But he still has the gto he was buiding during the HPP days.

https://www.youtube.com/watch?v=7Lv0jAD9PtE

[Dragncar]

Molnar gives you a rod bolt stretch chart on the back of their instructions to keep track of it. Post #3 was very telling on just using a torque wrench. + - 25% is terrible accuracy. Proves Tom is right on this.
Another thing about Molars ARP 2000 bolts is that they are propitiatory to Molnar. No one else can use them. There is actually a slight thread interference fit when hand tight. The threads do not perfectly engage. But when the bolt is stretched to the proper length either by using a stretch gauge or by the torque + angle method the threads line up perfectly, full engagement. Tom thought of everything it seems. Nothing left to chance.

This is the 31$ tool i used for the torque + angle method. I have tightened for more bolts than the average guy in my life, and I am telling you the 30lbs + 60 deg was a lot more than the 75lbs torque many 7/16 rod bolts use.
https://www.amazon.com/gp/product/B0...?ie=UTF8&psc=1

My buddy who has every engine building tool in the book used the stretch method and the torque it took to get it was 92lbs ! He was the one who suggested I use the torque + angle method.

This thread is about a Chevy guy who used Molnar rods and crank on his 1002.7 HP 582ci pump gas BBC build . Dyno sheet is on page 3. Screenshot would not enlarge.
https://www.chevelles.com/threads/58...#post-10912784

[Gach]

This all makes me wonder my BME aluminum rods required no stretch gauge to torque them. He give me a specific torque number. They Have 30 dyno pulls about 20 quarter mile passes and a ton of street miles

[Schurkey]

Originally Posted by JSchmitz

I have an old Craftsman clicker. You should also back them off to zero for storage.

No.

Micrometer-style torque wrenches should be turned to the lowest torque setting on the tool, not to "0".

For example, a 50--250 ft/lb torque wrench would be turned down to 50 ft/lbs when you're done using it, not turned down to "zero".

"Split-beam" torque wrenches are also "clickers" but they don't need to be turned down at all. They can be left with the torque set to any value.

Originally Posted by JSchmitz

Critical to use specified lube on specified surfaces. Make sure to coat the washers and threads. I think a lot of people coat the threads and forget the washers. The thrust surface under the nut is a major source of friction.

Absolutely true.

However, don't coat the underside of the washer that goes against the casting/forging. Coat only the side of the washer that bears against the fastener.

Coating the casting- or forging-side of the washer can lead to over-tightening the fastener.

[Dragncar]

I think if you are using the stretch only method it would not matter if you had grease on both sides of the washer. Stretch is stretch, In torquing the bolt to 30lbs then 60 deg it could effect it a little bit.
Torque wrench only, even more so.
It is kind of tough to get the grease only on one side, kind of messy.
Question, coarse threads vs fine threads like main studs vs stock bolts. Is 100lbs of fine thread torque really more than 100lbs on coarse threads. If so, how much ?
I was going through that with my played caps. Trying to replicate what Whitmore did when he line bored it, but he passed away so I just had to go with stock numbers on fine threads. Machinist side it was dead straight and round when he checked the mains. He was impressed.

[JSchmitz]

Originally Posted by Schurkey

No.

Micrometer-style torque wrenches should be turned to the lowest torque setting on the tool, not to "0".

For example, a 50--250 ft/lb torque wrench would be turned down to 50 ft/lbs when you're done using it, not turned down to "zero".

Ok. I learned in school to turn it "all the way down". Maybe it was supposed to be lowest setting. That was a really long time ago. Lol! It makes more sense to me to take all of the load off of the spring. Thanks for the info.

[mgarblik]

Originally Posted by Gach

This all makes me wonder my BME aluminum rods required no stretch gauge to torque them. He give me a specific torque number. They Have 30 dyno pulls about 20 quarter mile passes and a ton of street miles

I think I can answer this one. As you know Bill Miller comes from the top fuel world. His primary focus is nitro engines and other "real" race engines as he calls them. He specifies motor oil only on his connecting rod fasteners. No special sauce from anyone. Rod stretch tools and even torque to yield is impractical when the rods are going on and off hot in an oily environment in a 75 minute time interval with multiple people handling the parts. His bolts and the threaded area in the aluminum rods are way over-designed for a given application. His "torque value" puts the fastener in a very wide elastic range of the bolt. So even if the torque value ends up +- say 15-20%, it is still in the elastic range and the clamp load is sufficient to keep the joint closed.

As far as Molnar rods having some kind of interference fit between bolt and threaded hole? It's possible Molnar has some sort of special machining process that produces a slightly higher engagement % than the standard 75%-up to 85% on a precision fastener set. Or he has ARP spec. a slightly larger fastener blank to produce the same higher %. Actual interference would create friction which is the enemy of the joint and a wild card when tightening. Next time i talk to him at PRI, I will ask. When I run the bolts into his rods, I don't feel any resistance until they seat.

[Elarson]

Originally Posted by Shiny

This is not obvious to me. I guess I expected more of a "free body diagram" exercise where the bolts actually compressed the rods to some degree. In other words, the rods and caps are loaded and compress to counter the bolt expanding. What I'm struggling with is why the stress acting on the rod (ie inertial loading from piston and rod mass) doesn't add stress to the bolt? I get that the inertial loading would have to exceed the clamping force to cause complete separation of the cap, but wouldn't inertial force cause the bolt to elongate since it adds to the clamping stress? I picture the bolt elongating and the compression force within the rod and cap being reduced, so they also stretch.

You're so very close to the answer! When initially preloaded, the bolt face to cap face has equal and opposite load. The cap parting line has equal and opposite load. So everything is in equilibrium. For illustration, lets pretend that the fastener preload and parting line load is 1000 lbs.

Now if you apply 200 lbs of external loading to the cap, the bolt load does in fact see that extra 200 lbs. But....the cap parting line load is reduced by 200 lbs. So now you have 800 lbs clamp load at the parting line plus 200 lbs external load and the fastener gets the sum of the two (800+200) and still is at 1000 lbs.

If the external load is right at 1000 lbs, the cap parting line has zero load and the bolts sees 1000 lbs external load plus zero load from the parting line = 1000 lbs.

If external load is 1500 lbs, the bolt gets 1500 lbs of load from that and since the cap parting line is open and waving in the wind, it doesn't contribute any load to the fastener; so the fastener gets 1500 lbs + zero lbs = 1500 lbs.

So bolt load doesn't exceed preload until you totally unload the parting line joint.

As a side, we've done detailed FEA analysis of bolted joints and the actual fastener load ends up about 2-3% higher than the above discussion.

Eric

[carcrazy]

Excellent information...thanks!

[Shiny]

Eric, thank you! I appreciate you taking the time to educate. Your explanation is perfect and I now "get it". This should be one of those interview questions like thermal expansion of a donut, as it's not intuitive, at least for me... and I worked with MechEs for decades mitigating fatigue risk!

After my last post, I tried to educate myself and found a couple useful sites online.

Wiki has this cartoon illustrating exactly what you just clarified:



ARP has more insight (link):

"Pre-Load

Finally, although not a design parameter, the subject of bolt installation preload must be addressed. It is a fundamental engineering concept that the force in a bolt in an ideal preloaded joint will remain equal to the preload until the externally applied force exceeds the preload. Then the force in the bolt will be equal to the external force. This means that fluctuating external forces will not cause fluctuating forces in a preloaded bolt as long as the preload exceeds the external force. The result is that fatigue failure will not occur.

In a non-ideal joint, such as in a connecting rod, the bolt will feel fluctuating stresses due to fluctuating rod distortions. These are additive to the preload, so that fatigue could result. In connecting rods, precise preloads are required because if they are too low, the external forces (the reciprocating weights) will exceed the preloads, thus causing fatigue. If they are too high, they provide a high mean stress that combines with the fluctuating stresses due to rod distortion. Again, fatigue is promoted. The objective, then, is to preload a bolt so that it just exceeds the external load, and no higher.

To sum up: both insufficient preloads and excessive preloads can lead to fatigue failures."

Great education, thanks all.

Mike

[Stan Weiss]

Originally Posted by Gach

Are all your rod bolts same length have you checked them.

https://youtu.be/0CgqeJ-V094

What have others who use a stretch gauge seen as far as difference in bolt length? And other than having to zero the gauge for ever bolt does that small variance matter?

Stan

[Gach]

Stan.I don’t really know but this video just makes one question it. Would it be something that would make me want to check every rod bolt length. I’ve never used rods that needed rod bolts stretch type of set up. Mgarbilk and Eric did an excellent job on helping us get more educated.

[Dragncar]

Originally Posted by mgarblik

I think I can answer this one. As you know Bill Miller comes from the top fuel world. His primary focus is nitro engines and other "real" race engines as he calls them. He specifies motor oil only on his connecting rod fasteners. No special sauce from anyone. Rod stretch tools and even torque to yield is impractical when the rods are going on and off hot in an oily environment in a 75 minute time interval with multiple people handling the parts. His bolts and the threaded area in the aluminum rods are way over-designed for a given application. His "torque value" puts the fastener in a very wide elastic range of the bolt. So even if the torque value ends up +- say 15-20%, it is still in the elastic range and the clamp load is sufficient to keep the joint closed.

As far as Molnar rods having some kind of interference fit between bolt and threaded hole? It's possible Molnar has some sort of special machining process that produces a slightly higher engagement % than the standard 75%-up to 85% on a precision fastener set. Or he has ARP spec. a slightly larger fastener blank to produce the same higher %. Actual interference would create friction which is the enemy of the joint and a wild card when tightening. Next time i talk to him at PRI, I will ask. When I run the bolts into his rods, I don't feel any resistance until they seat.

Interference fit was not the right word for their threads. Its been awhile since my talk with Tom.
What they are, asymmetrical threads. Or threads like a big C clamp would have. Flat on the top. If you put a Molnar 7/16" rod bolt in a Scat/Eagle rod there would be interference. They won't fit.
Standard "V" shaped threads have more load on the bottom/base of the thread and less on the tops.
The uneven loading is where problems/cranks start.
The Molnar thread results in a even load across the threads when tight.

[Gach]

The deciding factor for me would be listening someone with real experience and knowledge vs someone with very little experience . Plus what investment and what procedure I would be comfortable using.

[Dragncar]

My last post comes straight from the # 2 guy at Molnar, Ed. He is the one who sent me the 3 D cutout pics of the 3 rods.
Now seeing if we can get some 3 D printed cutouts of the PA and PA+ rods. They have a 3 D printer used for prototypes of rods.