Bolt failure thoughts hardcore engineers please!

I'd say you're right bar any lateral impact loading (the only thing you've not mentioned that springs to mind).

 

The big problem is fatigue - as the bolted connection is used the stress felt through the bolt will vary with the applied stress, whether from a wheel flexing in hard cornering or just applied torque. While the applied load might not overcome the pre-load, it is a cyclic stress. Over time that can fatigue. I saw a supercharged Camaro shear off all his lugs after multiple hard launches in an autox event. Wheel went flying, car dropped to the ground.

 

One factor you haven't included is that there might be some non-uniform heating going on -- so even though the TCE of the clamped materials and the bolt are probably pretty close (with them all being steels), there could still be some alternating tensile stress cycles in the bolt going on as things heat up (unevenly) and cool down. One other distinction I'd make is that I'd say that Verell's bolt "fractured" rather than "sheared", and it fractured where you might expect it would (i.e., at a stress concentration point where the cross-section changes radically with a fairly small radius). Additionally, the underside of the head is supposed to be perpendicular to the shank of the bolt, but everything has a tolerance, so if it's a little off this puts a big localized tensile stress at that small radius. And finally, the heat treatment is tricky/touchy where they want a very strong bolt, but not too brittle -- IME, it's not uncommon for them to mess this up and get a part that is more brittle than they really wanted. Just some musings...

 

Well the idea is that IF torqued to the specified amount and assuming the bolt size is correct, it should handle 99.999% of the fatigue stresses applied to it. Yeah I'm not about to get into those calculations, waaay too many unknowns. Just torque to the spec'd #'s and go with it

 

Just to clarify and see if I understand the question correct. When you tourqed the bolt and left everything static ( no load on the drive train etc) the bolt is only in tension with no shear load. When you put a load on the drive train the bolt can fail in three ways in this model, it can be sheared if one flange turns in relation to the other or the bolt could be pulled apart in tension or it could be strained repeatadly and exceed the fatigue limit. I'm not sure which one you are asking about? Any of these can happen if the other bolts are loose. The bolt can be pushed passed it's fatigue limit through repeated strain. If the other bolts are loose or missing and you put a load that would do something similar to prying one side of the flange apart over and over again straining the bolt it will evenutally reach it's fatigue limit and fail under a load much lower than bolt was designed to handle.

The amount of friction that the two flanges would have to overcome to put a shear load on the bolts would depend on the finish and type of material as well as the clamping load from the torqued bolts. It would be fairly easy to overcome the clamping friction if the surfaces where polished and hardend and any lubricant on the mating surface on assembly would make it even easier.

If you drive a car around with one lugnut you might go many miles before it fails regardles of if you maintain the same speed on flat roadway. Even though the load never exceeds what the bolt can handle evenually it will fail from repeat strain when it reaches the bolts fatigue limit. The bolt will appear to have sheared off but really it just failed in tension after being bent and streached(strained) over and over again.

 

Yes, if all of the other bolts are still tight, and there's no evidence of the ring gear moving relative to its mounting, I think you have to conclude "bad part" or "improperly installed part".

 

Not an engineer and did not sleep at holiday inn express but I have seen this happen 1 break others tight. I think as the above person "snip...bolt can be pushed passed it's fatigue limit through repeated strain." Also, In my hillbilly way of thinking often removal torque is greater than original application torque on just a generic fastener that has sit for a while. So it also stand to reason that you can have some minor maybe not detectable or within spec about of torque on the other bolts but yet actually had movement and loss of cyclic clamping load thus failure of a single bolt. The rest just did not break yet.

 

He's correct. Friction between surfaces is one factor in the calculations of the number of bolts needed, thread size, tightening torque. Basically if the two flanges only have to bear torque load the bolts are calculated for a tightening torque of 90% limit of elasticy. If there's additional longitudinal load, this has to be considered in the calculations because this load adds to the bolt's pre-stress.
And it's difficult to calculate the dynamic load which will break one fastener when there's any sheer applied to the bolt. Not sheer is the problem, but bending stress which then occurs on the bolt and which is basically forbidden on regular threaded bolts.
One could try a bending calculation but based on my experiences one will get results with very LOW strength depending on shape.

On serious heavy machinery applications we regularly renew any bolts which haven't failed yet, but have seen bending stress because they were loose.

Best Regards from Germany

Martin

 

I think what your trying to discuss here, is the clamping load the bolts provide between the carrier and crown wheel, the friction provided, and how it reduces the shear loads on the bolts? Obviously whatever is the weakest link in the chain will yield if the forces applied exceed its ability, and some are a bit unknown, so the manufacturers generally overbuild and dont have any trouble. I personally feel that Verells broken bolt was a total fluke. Many people have tracked these cars over the last 30 years, and the differential has not been a real issue. But ultimately I think once the crown to carrier friction is overcome, the bolts themselves are really going to be carrying the brunt force of the shear load and for all itents, the clamping force/crown to carrier friction provided by the bolts, is irrelevant.

I feel that if you can reduce slop by tightening up all the clearance's in the carrier/spider gears, youll relieve a great deal of stress and will be able to sleep at night. But while you have it apart, send the bolts out to be x-rayed by an aviation shop. They use bolts/studs to attach propellors to the crankshaft flange of aircraft engines, wing bolts, engine mount bolts, etc., those guys are very well trained to see things that would get past virtually anyone in the automotive world.

 

isn't it the hub centering which supports the side load on the wheels ? I see a very precise fit between my hubs and rims and this should be taken into account when using any spacers or aftermarket wheels which DO NOT have a perfect fit regarding the hub center.
Some folks think, that the wheel bolts center the wheel, but this is not, how things should be.

Best Regards from Germany

Martin

 

o.k., I thought with side load you meant lateral load; the load perpendicular to the bolt's centerline.

Best Regards

Martin

 

IMO, Yes- if you are talking about bolts used to clamp 2 pieces together.

 

No, that is not necessarily a true assumption. You are mixing two different load conditions --- tension and shear.

Torqueing a fastener to clamp a jointed connection preloads the (bolt) in tension, of course. If the design is accurate (i.e., proper number of bolts, proper size / material, and proper preload), the tensile stress in each bolt will not increase to critical when the clamped joint is loaded throughout its normal operation.

It is standard in joint design to torque fasteners appropriately in order to stress them to between 50% and 90% of their yield strength to ensure that sufficient preload is maintained to prevent gapping and slipping of the joint. As you mention, if slipping occurs, the fasteners are then exposed to shear loads, which combine with the pre-existing tensile loads and can cause the bolt to fail.

It is never a good idea to let threaded fasteners take additional shear loads as the stress concentrations at the thread roots can easily become critical and lead to failure.

Loss of design preload can allow shear to come in to play. Two main scenarios cause this ---- (1) under-torqued fasteners and (2) a joint that undergoes dramatic operational temperature changes.

When joints are assembled at ambient temperature and then go colder --- preload is typically reduced. In Aerospace applications, this is nearly always the case, and we avoid shear loads by preloading bolts to 90% of their yield strength and putting in shear features when needed.

However, in automotive applications the opposite case occurs --- joints are assembled at ambient temperature and then typically go warmer which will generally increase preload. So, if a fastener is over-torqued, and then the joint gets warmer, the allowable tensile stress in the fastener can be exceeded to failure --- this happens often in the automotive repair world as some folks tend to "over crank" things with the idea that if "tight is good, tighter is better". Conversely, if fasteners are under-torqued, slippage can occur, and shear loads enter into the mix, which can also induce failure ---- either directly or by combining with the tensile loads.

Without personally examining the particular failed components that you are dealing with, I can't make a certain determination as to which failure scenario has occurred. But, I'm going to make a fairly safe bet that the Engineers at Ferrari designed the joint correctly ----- so, somebody likely didn't torque a fastener (or fasteners) to spec.

However, even Ferrari makes mistakes and not everything that left their design tables is immune to "improvement". And, I know from reading your posts that you are someone who understands how and is able to successfully modify parts .

So, if you want to eliminate the possibility of this ever happening again, I would suggest that you torque the fasteners to spec, of course, and add a few shear pins to the joint. Assuming that there is enough material in the mating areas of the components, adding 2 or 3 solid steel pins of 0.150" - 0.190" diameter would be plently adequate to take any potential shear loads (that's just off the top of my head ---- but, I'll stand behind that SWAG ).

 

???

By the time you have actual bending of a fastener --- even in terms of short beam theory type ---- you are done with a clamped joint design. Gapping is easily avoided with correct joint design (number of, size / material of, and pre-load of fasteners) ---- also, providing that is installed and maintained correctly.

Excessive tension (due to thermal excursion or over-torque) and critical loss of pre-load allowing direct shear due to slip are the only failure mechanisms in an otherwise properly designed clamped joint. Gapping sufficient to induce fastener bending falls into the "who f!@#ed up!?" realm of design or installation / maintenance.

 

YIKES!

...pins, baby, pins!!!

Good luck with the rebuild.

 

MARTIN ---

I really like the picture you have posted in your profile Is kitty a lead mechanic or just an apprentice?? --- LOL

Cheers