electronic torque wrench

Seems like a good idea for checking the torque wrench.
Would at least get you in the ball park.
After all you can't deny facts
 
yep, my son's real fussy about torquing stuff, which is good and I have weights which have been certified the last time they were used. It would be a fun experiment to calibrate his torque wrenches with weights used to calibrate scales.

... I think to do it right I'll have to wait till it's warmer. ... :(

ps... lol, I calibrate my $10 scales with $2000 weights.
 
I have a couple of electronic torque wrenches and the MetCal paperwork that goes with them; they are amazingly accurate. That is the funnest part in the end, as of all the things that effect preload in a bolt or stud and all the things that effect fastener to fastener preload variation around a bolt circle (there are about 76 different items, a fact not widely known), torque wrench accuracy is about the least important and has the least effect when the same torque is applied to each fastener by the same wrench by the same technician, except in the very unusual case that the torque wrench is truely defective. Basically, a torque wrench is the most common way to assure that the same tightening force is applied to each fastener in any bolt group, and that force can be witnessed by an inspector as it is applied if it is a "critical" joint requiring that, and, whether witnessing installation by an inspector is required or not, it yields a number that can be recorded on a joint assembly record sheet for quality assurance purposes.

What it does NOT do, regardless of torque wrench accuracy in a properly functioning torque wrench, is assure that each fastener in a bolt circle ends up with the SAME preload as it's brothers with any of the precision most people believe. For example, in a statistically valid sample (and that can be as few as 10 fasteners) using a torque calculated to produce, say, 2/3 of yield in each bolt or stud, when you are done, all you really know is that each fastener in the bolt circle, which you so carefully tightened to that specific torque, has ended up with a preload equal to somewhere between 40% and 90% of yield. That's quite a range, considering you used a wrench with a guaranteed accuracy of 4% or less (a LOT less in tha case of an electronic torque wrench)! There are standardized tightening procedure used in many industries that, if followed properly, can skew this more or less bell shaped curve of preload distribution in a favorable manner; they don't change the high end of the distribution, but thay take some of what would have been the low end, those between 40% and 55% of yield or so, and fold them into the middle, so that you don't have as many fasteners under 2/3 of yield (or whatever the target figure was; depending on the fastener material and the joint design and intended service, that value can commonly be anywhere from 50% of yield to 95% of yield), or at least they aren't nearly as far under 2/3 of yield as 40%. But the bottom line is, people are assuming that since they are using a quite accurate tool, they are getting a product that yields results a lot better than a range of -27% of the target value to +23% of the target. Unless lady luck is really smiling on you that day, an experienced technician using an approved procedure with a torque wrench with a current calibration sticker can only hope for a fastener to fastener preload range (from highest to lowest) around a bolt circle of 30-32% on the best day of his life. And, for a little more humor, that same EXPERIENCED technician can hit 35% variation around the bolt circle using mechanic's feel and a box end or regular socket wrench - but mechanic's feel doesn't yield a varifiable number to prove he met a specified preload spec, and a torque wrench does.

Most of this astonishingly large range of fastener to fastener preload variation is due to fastener to fastener friction coefficient variations, about which several chapters in whole books have been written; I won't try to summarize it here, but it gets back to those 76 factors influencing preload variation in an outline format list of such factors, and it is an eyeopener. But take heart; the people that designed that cylinder head joint or whatever are well aware of these variations and have taken them into account in the joint design and specified joint makeup procedures; bolted mechanical joint designs are amazingly tolerant of such things because they are, among other things, usually quite conservative with respect to material choice and size and number of fasteners.

And the economic solution, if you have the numbers available or a way to calculate the equivalent to a specific torque, is to use angular turn; fastener to fastener to fastener preload variation is on the order of 15% with angular turn, and you can achieve that accuracy with something as crude as a slugging wrench if the fastener is large enough and the target preload high enough to make that method of applying force the best choice. The beauty of angular turn is that it takes friction out of the equation, eliminating a whole bunch of those 76 variables.
 
So goodantight is close enough then.
Interesting, seeing as most of the newer head torques are in degrees of turn.
Done some structural work, specs call for as tight as a man can get it then 2 more turns. Now that's real close:)
 
I use them and like them Gear wrench is what Mac,Snapon and Matco are copying from you can get the electic torque wrench with degrees built in as well if you what 250ftlb plus 90 degrees set the wrench
 
flattop1 said:
So goodantight is close enough then.
Interesting, seeing as most of the newer head torques are in degrees of turn.
Done some structural work, specs call for as tight as a man can get it then 2 more turns. Now that's real close:)
Click to expand...

That sounds like fairly long fasteners, but it's typical in structural steel work - they use steel alloys in the same general range as Grade 5 steel fasteners and strive for preloads that are often in the 95-100% of yield range. It makes a great joint for the long term service and they can get away with that for two reasons - the resiliency of the alloys they use (percent elongation) and the fact that they aren't alloys subject to hydorgen embrittlement like Grade 8 and stronger steel alloys are. The same range of preload variation around the bolt pattern applies to those joints, but as long as the high end stays below 135% of yield, they just stretch and very quickly relax (the phenomonon is called short term relaxation, as the bulk of it occurs in the first 2-10 minutes after completion of joint makeup and the rest in well under 24 hours) back to 95-98% of yield range. High strength steels, grade 8 and stronger, with minimum yield strengths in excess of 130 KSI, won't survive that generally; if a lot of them don't break during installation because they are well over the target load, they will stand a high probability of failing in the long term due to hydrogen embrittlement. And since those structural joints are tapping at the door of problems, they use angular turn because it gives less fastener to fastener preload variation. Most of the joints I have dealt with use relatively short fasteners, and after "snug" (about 10% of the preload that would be applied with a torque wrench), they are tightened in three or 4 stages to somewhere in the 40-120 degree range, depending on the material and the fastener length.

One of the neatest applications of angular turn is for use and reuse of fasteners in repair of railroad tracks - not always the cleanest working environment around, and not the easiest to get the fasteners really clean prior to installation. Well, they figured out how much angular turn they needed, so all that was left was to define "snug" in a manner easily duplicated in the field without use of special tooling. Turned out to be easy, since for speed, they started the fasteners with a pneumatic impact wrench. Remember how the impact wrench goes Whirrr-toc-toc-toc-toc? They simply defined "snug", the starting point for the angular turn, as stopping the impact wrench at the very first "toc". Simple, quick, and consistant enough - a lot more cinsistant than a torque wrench would have been.
 
Since friction effects how accurately you can torque is there an ideal length of thread for the diameter of what's being torqued?

And can you also reference the ideal length of thread, per size, per type of metal?

And if there is such a chart, where can I get one off the web?

Here's my thought on it. Heads for instance only need to hold x amount of compression and water passages per the gasket. Let's say your putting cast iron to cast iron. Is there a length of thread and diameter that is considered sufficient and larger is not worth it or needed. How many of you need would be up to the engineers per the design. I can see if you had more fasteners a smaller size and thread length may be ok. ... ?

I found bolt size torque charts. I guess different size bolts must be torqued to an 'at least' amount or they will be prone to loosen because of not being stretched properly. I guess and just a guess on that, it's really a question.

But how much thread do you need to allow for stretch if it's needed. I don't see a bolt being stretched beyond a couple of threads in the hole. Wouldn't the hole also have to stretch for the bolt screwed into the hole to stretch?

edit: oooops found this, guess I'm in for some reading again ...

http://www.pcbloadtorque.com/pdfs/engineering fundamentals.pdf

edit again: took a quick look through and am I correct you might want to lube strong metal being torqued but never soft metal. I'm thinking if you lube thread going into aluminum, the torque necessary to clamp what your fastening, might be more then the aluminum can handle. ????? correct thought maybe ? and so NEVER lube a bolt going into aluminum ?????

edit again and again: Just had another thought on torquing. Lets say your attaching a head. I'm thinking you should take a reading of how much effort it takes to run the bolt in. Thread hole and bolt combo's which have less then perfect threads or maybe some debree in the hole will take more effort to run down. I'm thinking the bolts which took more effort to run down to have the same clamping ability would ... maybe... need less total torque ? ... or would they need a little more ? I think if there's more run down friction the torque will be less. any engineers out there want to answer it... ?

I'll try an answer... there's proly a minimum amount of stretch required for proper holding after clamping. And each bolt size has a range of torque verses stretch. So you have to have in most any condition just enough torque to create the minimum amount of stretch. Any more would be a waist of effort. ... but if there's not enough friction is the minimum stretch, if there is such a thing... enough ? darn how come such simple things are so complicated ? but there fun to thunk about. ... :)
 
so... if you want to increase clamping power holding a head on to seal both combustion and water, clean and lube threads to reduce friction, and clean and lube the head to reduce friction. ... I guess for a given torque, any reduction of friction increased how well things are clamped together. and ?

... sort of looks like if you had say, three different torque sticks it would be as good as any torque wrench, torquing in three steps.
 
Well, we'll try to answer these witout writing a book, Paul. Incidently, if you can find a book titled "The Design and Behavior of Bolted Joints" by John H. Bickford (CRC Press) in the library, it can answer just about any question you can come up with; any edition after the first edition even references a couple of friction coefficient projects of mine in the reference list for Chapter 7 (Torque Control of Preload). I'd recommend you buy a copy, but it runs around $125 last I checked (4th edition).

First, there is no ideal length for thread engagement when torquing fasteners, just the caution that you don't use a thread engagement below the minimum engagement arrived at by calculations. With a simple nut and bolt, IF you aren't playing games with materials, the height dimension of the nut takes care of that for you. With a set stud and nut type joint, there is the necessity for someone to calculate the minimum length of engagement and assure that the set end engagement is over that value. A rule of thumb for most "normal" material combinations (does not work well for Grade 8 or higher strength steel head studs into an aluminum engine block) is that the set end length of engagement must be at least equal to the fastener nominal diameter. It's all done with computer programs that reduce the calculations to a couple minutes (to get all the stresses in the different parts of the joint, including bearing stress of the nut and the bolt head against the clamped material takes about 8 hours of number runching by hand, about 5 hours after you've done a few). The computer is about .2 seconds after about 2 minute's worth of inputs, then you have to evaluate the results and do another iteration if you're not where you need to be. In your example of a steel engine block and steel heads, the thread engagement at the set end for the set stud or the bolt into the blind hole, an engagement of 1 diameter would generally be adequate, though the actual engagement would probably be more - conservative design works in the long haul. Also worthy of note - nuts are stronger than bolts/studs, as are blind holes, so you can often get maximum allowable stress on a bolt or stud with a nut or set end material that is not as strong as the male fastener. Also, heavy hex nuts (being wider in the across the flats dimension) are stronger than regular hex nuts, and also have a larger bearing stress area, so they stress the clamped material less at the same torque, which also means they can be used at a higher torque if bearing stress under the nut is the limit, rather than the nut or bolt strength,

Yes, high initial preload and minimizing fastener to fastener preload variation are the two keys to establishing and maintaining joint integrity, and that involves exceeding some minimum value, which is where generic torque tables can sometimes lead one astray. When establishing the required preload for a joint, some target figure is usually used, expressed as a % of yield. The thing is, it ends up (when you use a properly designed calculation regimen) being X% of yield of the weakest element in the joint - the bolt or stud, the nut, and in the case of a set stud or a bolt driven into a blind hole, the set end material, OR 150% of yield of the clamped material for bearing stress. Depending on the materials you are dealing with, any one of these limits establishes your preload - for a "flat plate" type joint. And that assumes the joint design was produced with the proper size and number of fasteners to begin with. Then there are curve balls, like raised face flanges in piping systems, where, depending on the pressure class of the flanges, you may find that the preload limit is nowhere near what the fasteners can carry as the fastener load limits would cause bending of the flanges. Or like gasketed joints, where the driver is usually the desired (or maximum allowable) gasket compression.

Permissable/allowable bolt stretch is a function of fastener length for any chosen diameter, regardless of what % of the length of the fastener is threaded; longer fasteners can be stretched more than shorter ones, which is why the target preload is usually expressed as a % of yield rather than as an increase in length due to fastener stretch. HOWEVER, IF you have the proper equipment AND unimpeded access to both ends of the fastener, a stretch keyed to that diameter, length, and material of fastener can be calculated for that preload, and stretch measured to establish preload. It's expensive and labor intensive and highly dependent on the skill of the technician, but it is an excellent method sometimes used in assembly of the crankshaft end of connecting rods, for instance.

Lubing the threads is desirable in most joints; whether the threaded fastener goes into a softer/weaker material isn't important - that gets back to establishing the target preload based on X% of yield strength of the weakest element of the joint - the torque to be applied would be limited to the yield strength of the aluminum the the Grade 8 fastener was driven into, as an example. What is important is knowing what thread lube was used as the basis for establishing the torque, especially if you are using a generic torque table - if the supplied torque was based on dry threads and you use a Molybdenum Disulfide based thread lube (Molykote 77 of GN, say), you would need to reduce the listed torque by over 50% or you would drastically overstress the weakest element in the joint because the friction coeficient of MoS2 lubed fastener is drastically less than dry threads.

What you are talking about in relation to the effort required to drive a threaded fastener into a threaded hole or threaded nut onto it is called running torque. In a normal situation, it is ignored. If the threads are so dirty or buggered up that real resistance is a problem, they need to be further cleaned or chased with the appropriate tap or die. If that isn't possible, you can measure the running torque with a torque wrench and add it to the specified torque. Where the joint uses nylock self locking nuts, the running torque should be added to a specified torque from a table for fasteners smaller than 5/8" - beyond that, the running torque is too small a fraction to worry about, but there is no reason not to add it if it makes the installer feel good, it's just extra work to determine the running torque. A way around that is there are specifications for reuse of nylock self locking fasteners which list a MINIMUM acceptable running torque. Just assume that value and don't worry about induced error - again, they're down in the weeds of fastener to fastener preload variation and don't matter.

You want to have more than just the minimum amount of torque necessary to get into the minimum stretch - we'd write another chapter if we get into short term preload loss, long term preload loss and fatigue issues - suffice it to say that when you are given a torque specification for a specific joint, it was arrived at with the intent to give both satisfiactory initial clamping load AND long term maintenance of sufficient load; it is well over the minum that would be required to make the joint work in the short term.

At installation, definitely clean both male fasteners and nuts or blind holes, then lube with the specified lube - if you reduce the friction coefficient too much, you run a risk of overstressing some part of the joint or overcompressing a gasket. If the torque specified is a dry torque, you can use motor oil or axle/bearing grease (NOT grease with added Molybdenum Disulfide as found in some bearing greases, which will announce that presence proudly - teflon additives are OK) as a thread lube and use the dry torque without any change (I have about a quarter of a million $$ worth of friction coefficient research a previous employer was kind enough to fund over about a 9 year period to prove that one). Thread lubes that actually reduce the dry lube friction coefficient way down there come from the anti-seize family of compounds; the pressure additives in motor oils and bearing greases are for a different type of pressure loading and don't make things any slipprier once you start loading the threads significantly.

If you would like some further insights into the quirks of bolted joints, check out this article in ARC Racing's web site (disclaimer - I wrote it :) ).
http://arcracing.blogspot.com/2005/04/on-rod-bolts-by-mike-gifford.html
 
thank you. We have been using gray anti seize in aluminum to help stop the reaction mating of the aluminum and steel, which we figure makes it hard to take steel bolts out of aluminum. Am I reading you correctly we should stop doing it because it will lead to over tightening and possible poor gasket sealing because of over compressing gaskets ? And to use motor oil or regular grease instead ?

it ain't real good when you don't get a proper gasket seal and hydrolic a rod.

thank you very much for your information
 
Definitely DO keep using the gray anti seize in the aluminum; wherever the toque(s) you are using came from, they fall in the acceptable range - since they are working fine, no need to change now, as anti seize is ALWAYS a better thread lube than any motor oil/bearing grease/general purpose lithium grease (with or without teflon added). One of the reasons to use a thread lube is to decrease fastener to fastener preload variation, as among the fasteners in a cylinder head joint. The biggest fastener to faster preload variations occur when threads are assembled dry. Since motor oils and bearing greases don't greatly change the friction coefficient in a bolted joint where it matters (at loads higher than running torque as you hand tighten and get into initial snugging down, before you grab the torque wrench and get serious) fasteners lubed with them demonstrate the same very large fastener to fastener preload variations - the only thing oils and greases do is give you that corrosion protection to keep the fasteners from fusing due to rust or other types of corrosion. With anti seize, you get the same corrosion protection and the highly desirable bonuses of reduction of fastener to fastener preload variation and another thing that oils and bearing greases won't give you - protection against galling.

Now, a dirty little secret for bolted joint assembly and achieving the right preloads - there is often a big difference between technically correct preload and technically acceptable preload. Here is a somewhat terrifying, but technically acceptable acceptance standard for bolted joints, especially applicable if you think you have overtorqued the joint - a warning, however; at the high end of preloads, use this standard only on joints with alloy steel fasteners, and use it with caution for steel fasteners at or beyond the strength levels of Grade 8 steel fasteners (more on that below). If the joint doesn't leak (where that is a factor), the fasteners don't break/strip at installation, and the joint remains tight in service for a reasonable service life, your preload is acceptable - maybe ugly on either the high or low side of desired preload, but acceptable until the next time the joint is disassembled for maintenance or repair reasons. This standard is usually applied when there is fear of a dramatic overtorque situation or a way out of spec undertorque sitiation (someone used the wrong table for instance, or the wrong column with the right table; torque tables are, among other things, limited to the values calculated for a specific material combination for the bolt or stud and the nut and/or set end) and the joint is now inaccessible due to the stuff assembled on top of it - not usually a problem in race cars. One factor in why this works is short term preload loss, which is discussed in the article on ARC's website I cited in post #10 above.

The exception to the above standard is where the fasteners are high strength steel alloys (minimum yield strength of 130KSI or greater, so grade 8 steel is the start of this problem); high strength steel alloys are subject to a phenomenon called Stress Corrosion Cracking, which can eventually lead to brittle failure due to Hydrogen Embrittlement. If the applied preload is above 80% of yield (a good figure for threshold stress level for embrittlement prone steel alloys; it can vary for other embrittlement prone materials), you stand a good chance of fastener failure due to H2 embrittlement. That's the bad news. The good news is that this is not usually a major problem outside of the marine environment (defined as anywhere from 50 miles inland on out to sea for H2 embrittlement discussions), though it can still result in random failure of a high strength fastener in a joint if all the necessary conditions are present for at least that one fastener. It's like the Fire Triangle we learned about in elementary school - remove any leg and no fire. In the case of SCC and H2 embrittlement you need an electrical potential (like when a joint has disimilar metals like a steel bolt or stud into an aluminum block), an electrolyte (water will do, salt water is even better), free hydrogen (gonna' have that wherever you have an electrolyte) and high stress in the fastener(s), defined as stress above the threshold stress level. H2 embrittlement is not generally a factor in high strength steel fastener failures in racing machinery simply because it is not usually wet for the few weeks to few months necessary for SCC to develop. It is never a factor in services like rod or main bearing capscrews inside a well lubed engine.

However, failure due to low cycle fatigue can occur in joints subject to cyclic loads (aha, back to our cylinder head bolts again, and rod bearing capscrews are another), and low cycle fatigue is the gremlin to watch for when preloads appear adequate but are too low. In the case of low cycle fatigue the cyclic load never needs to exceed the yield strength of the fastener, it just needs to exceed the load carrying ability of the joint at that preload. Not usually a problem on race cars and their components, especially where a specific torque is not specified and the general "ok, we'll just give it an extra tweak to make sure it's tight" philosophy of racers putting things back together is present. Where it gets to be a problem is if someone uses motor oil on rod bearing capscrews and the torque specified for an MoS2 or graphite based thread lube - that was the beginning of the email exchange between me and Tom at ARC Racing that ended up with the tech article on their website - their assembly instructions for their billit rods allowed for this, and not enough torquing instructions do. To show why it matters, typical mean friction coefficient for alloy steel bolts with alloy steel nuts is 0.07 with an MoS2 based thread lube, 0.11 for a graphite based thread lube and 0.19 (quite a jump!) for unlubed or lubed with oils or bearing greases rather than anti seize; basically if you have a torque based for sure on dry assembly and go to a graphite based anti seize such as what we usually purchase at NAPA/Advanced Auto/Auto Zone/ Ace Hardware, you need to reduce that torque by 40%-45% to achieve the same preload rather than overtorqueing a whole bunch.

Two final thoughts; one, it is not necessary to look for a procuct labeled as a thread lube - all the good ones are just regular old high temperature anti seize compounds, althought there are a couple out there labeled as thread lubes that are just anti-seize compounds repackaged and relabeled by the company selling them, or inhouse copies of high temp anti seize compounds. Anti seize compounds (mostly graphite based; the silver ones have also ground up microfine aluminum or zinc added, depending on the brand, and the ones that are the color of somewhat dirty pennies have ground up microfine copper added -either type is fine) can be had at any of the outlets listed at the end of the last paragraph. And two, NEVER assemble stainless steel fasteners without an anti seize compound type thread lube. SS bolts/studs and SS nuts are the second most galling prone combination we have ever tested (titanium on titanium is the worst); you may have been lucky up until now with SS on SS so don't push your luck - we have had SS nuts freeze up on bolts due to galling while hand tightening them in the torque test fixture!

Now you understand the comment I made in my first post about trying not to write a book.... :)
 
Thank you again.


... and me being me reading tech stuff always makes me have a question. ... :)

You wrote: :In the case of SCC and H2 embrittlement you need an electrical potential (like when a joint has dissimilar metals like a steel bolt or stud into an aluminum block), <snip>."

Spark plugs use high voltage. In a race car the block is often in the ground path. We race push to start stuff without a battery. Reading what you wrote I can see the need for a ground strap between the heads, the distributor and where the mag is mounted, bi passing the block.

It would eliminate any electric leg needed to cause embrittlemen, by providing a return path which does not include head bolts. Maybe over kill but an easy thing to do and it would not be a bad idea to help cylinders fire.

Thank you again


paul
 
I find this quite interesting.
the stainless steel is the worst I've seen for galling as you stated even by hand with lose fit threads.
Recently have been working with raised flange joints.
 
Glad you find it interesting flatttop; I had a steep learning curve when I first stumbled into the fastener world, and the learning didn't stop for the next 21 years, the curve just leveled off a lot, lol. Are the raised face flanges by any chance ANSI B16 flanges?

Paul, That isn't a bad idea at all, but unless we're racing in the coastal area of the marine environment , not really necessary, and probably not necessary there unless the car is stored outside, although it does (rarely) happen that the necessary conditions exist at the micro level (across the ends of a short surface scratch, for instance) for SCC and eventually H2 embrittlement to develop. Stray currents from installed/associated electrical systems are just as good a source of the electrical potential needed as the battery effect of dissimilar metals - I was once told only half in jest by another engineer that the whole world is a battery waiting to happen; after I got into the study of brittle failures caused by stress corrosion cracking developing into full blown hydrogen embrittlement, I realized that he was pretty much spot on. The main thing with an engine is that they get dried out regularly after washing, they have lubes on the inside and WD sprayed around them on the outside, plus firing them up drives moisture away really effectively. My advice with that idea is that if you ever start getting unexplained random failures of high strength steel fasteners (like ARP head bolts or studs) on an engine, wire it up as you've described, as it will probably make the problem go away. These failures don't always occur in service, fortunately; they are usually random in nature (well, now we know which fastener in that bolt circle was really well polished and overtorqued...) and where normal precautions are taken during joint makeup (don't overtorque, in other words), they generally do not occur at all. Because they are generally random in nature, the most common scenario for finding them is on joint reassembly with reused fasteners - one or two fasteners in a joint will be effected, and well before full preload is reached - on the first or second pass of a three or four pass plus check pass(es) installation procedure, that fastener or two will break, at which point one generally takes the fasteners back out and replaces them all just to be safe. But unless all of the conditions (electrical potential, an electrolyte, free hydrogen (usually from the electrolyte) and stress above the threshold stress level) are present for long periods of time (several weeks minimum if the item is wetted constantly), SCC and worse is unlikely. But another curveball encountered a number of years ago - hydrogen in the metal causing SCC doesn't always come from the environment. Plated fasteners (such as cadmium plated fasteners) can fail if they have not been been properly heat treated after the plating process to drive off residual H2 absorbed by the steel during the plating process - that's a manufacturing defect that the Department of Defense encountered in the mid or late 1980s. And I've never seen a failure that started out with SCC in high strength internal engine bolting like rod or main bearing fasteners - lots of motor oil and elevated temperatures in that environment seem to prevent it. I have seen internal bolting (an OEM rod bearing cap screw) fail due to improper inspection prior to assembly - the engine builder magnafluxed the bolt in the rod rather than pressing it out it and doing the whole bolt - by doing so he saved us $6.50 per rod for a quick cleanup/clearancing of the big end of the rod after pressing the bolts back in. The bolt in question failed at the junction of the shank and the head, a common failure area of bolts stemming from the occasional manufacturing defect in the fillet area of the transition. The bolt failed about 200 feet before the traps on the first ( and only) day in the life of that engine - we still got the win light on that pass (the explosion was so impressive the guy in the other lane backed off in case we swerved across the dividing line), but we had to scratch for the next round, lol. We also found an engine builder that knew how to inspect new OEM bolts properly for defects in the most likely failure areas prior to engine assembly.
 
Yes b16 Flanges.. propane and butane spheres. lin/lox and some ammonia tanks Even on the water towers. They seldom under design anything.
The bolts are A325 and A307 . Not many get the 2 full turns more, I think the engineer just wants to be certain there tight.
 
We used those flanges on all kinds of systems. Like you said, they aren't under designed, lol. All our services used spiral wound gaskets (Flexatalic was the most common brand), and the gaskets used the same pressure rating system as the flanges, so for instance, you use a Class 600 gasket for a class 600 flange. One reason that this was important is that the flanges are so beefy that once the design was finalized years ago, they essentially ignored the flange strength when establishing the fastener torques for joint makup - the torques are based on establishing adequate compression of the spiral wound gasket. Any preload in the adequate range for that won't overstress the flange. But you can't use a Class 600 gasket with Class 300 flanges in a pinch - the preloads you can achieve with the fasteners in a class 300 flange won't adequately compress the highter pressure class flange, so you will get leakage. On the other hand, the design of those raised face flange joints is, theoretically, leak tight at essentially zero fastener preload, and my division funded a series of tests for another reason that, as a byproduct, proved this was true. The test fixtures were set up with class 600 flanges and flange gaskets that were within spec for compressibility and gaskets that were out of spec on the hard side (the reason for the test was manufacturing compressibility being out of spec on the hard side from one manufacturer, NOT Flexatalic, incidently) and all of the joints with gaskets within spec passed, holding 600psi steam at slightly over finger tight. Even the out of spec hard batch held pressure down to about 8-10% of their required torque. Of course all of that performance is degraded if careful attention isn't paid to flange alignment during system fabrication.....
 
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