Are higher grade bolts more brittle?

Unless you are 15000 feet in the air when something goes wrong. Rather bend and hold than snap. You can't afford the extra weight to make something that will neither bend nor snap.

Reply to
clare at snyder.on.ca
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As was pointed out here, at the lad at which the low grade bolt would completely fail, the high grade bolt would not even bend.

i
Reply to
Ignoramus29897

Granted.

I guess I expressed it a little TOO simply (?)...

However, there are a lot of cases where the clamping force is minimal and the bolt is simply a shear pin.

Ultralight wing tube style spars - for one. Can't clamp much without crushing the tube.

Reply to
cavelamb himself

Thus very likely transfering an excess load to even more critical structure...

Reply to
cavelamb himself

When you are cork screwed into the round, does it really make a difference which part failed first??

I do understand the logic to a point over using Grade 5 versus Grade 8, but are we forgetting that it makes a difference what the design of the whole assembly is? I can see why a Grade 5 might be better than a Grade 8 in a assembly that was designed for the lower strength bolt. Same with an assembly designed for a Grade 8 over the Grade 5 bolt. Back the the OP's question, as to mounting a trailer hitch. The Grade 5 will more than likely do the job just fine, and the Grade 8 will give a larger margin of safety, probably over kill, but I surely would not worry about a Grade 8 bolt failing on a typical trailer hitch, something else will probably fail first, like the frame member itself! A few years back A friend of mine got rear ended, he was pulling a small boat with a Chevy pickup, and a Class III hitch. He was hit so hard that vehicle crushed his boat and rammed into the back of his pickup. Looking over the wreck latter we noticed the six, 1/2 diameter, Grade 5 bolts holding the hitch onto the frame were still tight, and appeared to be snug, even though the hitch and rear portion of the frame was a tangled mess! Greg

Reply to
Greg O

Have you ever built a plane? You need to take a REAL GOOD LOOK at the way they are designed, and figure out why. It just does not work the way you expect. There are VERY GOOD reasons why AN hardware (aerospace standard) are essentially grade 5 bolts with a pedigree.

Reply to
clare at snyder.on.ca

And VERY poor design when built that way. The proper way is to sleeve the hole so the bolt is supported for it's full length instead of just by the skin of the tube, and the bolt can be torqued to it's proper torque to provide maximum strength. Then they could likely even get away with lighter bolts.

Reply to
clare at snyder.on.ca

Grade 8 bolts have failed in trailer hitch mountings. - Where grade

5's would have held. (at least long enough to create a rattle to warn of impending doom) If a bolt is not providing almost exclusively clamping force, DO NOT use a grade 8 - particularly if there is impact loading or reversal of forces that will put impact shear loading on the bolt.

Looh at ANY light to medium duty hitch kit that fastens to either a unibody or a typical automotive frame. The bolts are grade 5 AT BEST - more often grade 2 or ungraded utility bolts.

Pickup frames and hitches are somewhat different in that the hitch brackets are USUALLY a very good fit to the frame rails, so a properly torqued bolt provides a LOT of friction between the frame and the hitch brackets. Enough that the bolts are not taking any appreciable shear loading.

Reply to
clare at snyder.on.ca

Can you name one UL manufacturer who does that?

Reply to
cavelamb himself

In the case of a seat belt anchor, the amount of energy it can absorb before finally breaking is going to be important. It depends on what you consider to be strength. I would consider strength to be a joint's ability to resist the forces applied to it. And so a stronger joint would be one which fails under a greater force. But you might also consider strength to be a joint's ability to absorb energy before breaking. Both are important, though sometimes one is more important than the other.

I disagree. Take a simple example: a single-shear joint made between two mild steel flat bars. There's a hole in each bar, and a bolt connecting the two holes.

Either the bars are weaker (the bolt will tear through one of the bars when the joint fails), or the bolt is weaker (the bolt will fail in shear). If a grade 5.6 bolt is weaker than the bars, then substituting a grade 8.8 bolt can only make the joint stronger. If a grade 5.6 bolt is stronger than the bars, subsituting a grade 8.8 bolt will make no difference.

Now if the joint is part of a large and complex structure, it's possible than using a weaker but more ductile bolt might, in some circumstances, impose a safer distribution of forces within an overloaded structure, making the whole structure stronger.

But even though the whole structure might be stronger, the individual joint is still weaker. I can't see how the individual joint can possibly be made stronger by substituting a weaker but more ductile bolt. If you disagree, please explain why. Perhaps we agree and it's just a misunderstanding?

A weaker structure, possibly, but the individual joints are still stronger.

Best wishes,

Chris

Reply to
Christopher Tidy

I think you'll find that if the hitch bolts are tightened to the correct torque, the load is carried by friction.

Chris

Reply to
Christopher Tidy

I'm inclined to agree with Nick here. The few times I've made this mistake myself in the past, the joint has come loose. Sure, there are a few instances in which it can be okay to load a bolt in shear (a shackle for example:

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but they're cases in which loosening isn't an issue. Mostly it's a bad idea.

Best wishes,

Chris

Reply to
Christopher Tidy

I think you just disagreed, and then agreed. d8-) Yes, it applies mostly to structures with multiple joints, or with multiple fasteners in one joint.

It would have to be a very contrived case to make the point with a single fastener, but the principle still applies. In a car or aircraft crash, for example, you could have an extreme overload applied to a joint, bot only through a distance of a fraction of an inch. If the bolt doesn't give, something will break.

I wouldn't get into energy because it complicates things, although it's an issue with impact. We're just talking about distribution of forces, or, in the case of a single fastener, an excessive force applied through a very short distance.

A joint can have more than one fastener, as they often do in bridges and vehicle structures.

-- Ed Huntress

Reply to
Ed Huntress

I did. But I've thought this through very carefully, and I think I'm prepared to mostly concede the argument. Suppose a joint has many bolts arranged in a line parallel to the force on the joint, and that all the bolt holes are equally spaced when the material is in the unstressed condition. The joint is then overloaded to the point where the limiting friction is exceeded. The first few bolts will carry most of the load, because the material between the bolt holes stretches, causing the other bolts to go slack. It's the same as the way in which the first five turns of a long screw thread carry most of the load. So if relatively brittle bolts are used, it's possible that the first bolts might break before they transfer enough load to the bolts further down the line, then the bolts further down the line will break, and so on until the joint fails completely. The same would be true for rivets.

But I suspect this is only important for joints with many bolts or rivets. Perhaps more than five in a line parallel to the force, at a guess. Of course, if the load on the joint was constant and known (not an overload) it might be possible to vary the spacing of the holes to ensure that all the bolts carried an equal load. Now there's an interesting idea.

Using a bolt in single-shear, perhaps. But a pin in double-shear or multiple-shear is pretty common. Think of a tractor's three-point linkage, a towing hitch or an eyebar suspension bridge.

The selt belt anchor is certainly a case in which energy is important. I just raised it because I wasn't absolutely certain that you weren't talking about energy at first.

I don't believe this argument applies to a joint consisting of a single bolt, considered in isolation.

Thanks for an interesting discussion. This is what I come to RCM for!

Best wishes,

Chris

Reply to
Christopher Tidy

Not off the top of my head, but I think the Kolb 500 trainer (2 seater) is sleeved.from the factory. A friend a few years ago had one that was sleeved Also I think the Beaver 2 seater does. I know a friend's Beaver does, might not have been from the factory. I know a lot of ultralights have SCARY engineering (or lack there-of) and many owners modify them to make them significantly safer.

By "sleeved" I don't meed doubled tubes. I mean the hole is drilled overside and a piece of tubing is welded or brazed into the hole, the right size to fit the bolt snuggly.

Many ultralightes also have holes drilled through the tubes the wrong way, seriously weakening the tube.

Reply to
clare at snyder.on.ca

Automotive seal belt anchors are GENERALLY designed in such a way that the floor pan will deform before the bolt or the belt breaks. The thin floor is doubled with a heavier plate that retains the extra strength bolts.

If the two bars are properly joined with a properly engineered joint, the bars themselves will almost stretch before either the bolt or hole deform.

Reply to
clare at snyder.on.ca

In a well engineered and properly installed hitch, yes.. In this case, the bolt is providing almost exclusively clamping force.

Believe me, there are LOTS of hitches out there where neither of these conditions is true.

Reply to
clare at snyder.on.ca

Since you've put some thought into this, you may find that reading a more sophisticated engineering treatment of it would be worth your while. It's too far in the past for me to remember much of it but there is plenty of material on ductility and brittle failure in the field of aerospace engineering. It's a very big issue there.

You could try the SAE website, which has a bookstore and white papers in the aerospace division.

-- Ed Huntress

Reply to
Ed Huntress

Pardon the messy line wrapping - just wanted to show what's there...

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Astronautic Structures Manual (On-Line), NASA MSFC (Marshall Space Flight Center), 1975. Astronautic Structures Manual, Volume 1, NASA TM-X-73305, NASA MSFC, August 1975, 839 pages. Astronautic Structures Manual, Volume 2, NASA TM-X-73306, NASA MSFC, August 1975, 974 pages. Astronautic Structures Manual, Volume 3, NASA TM-X-73307, NASA MSFC, August 1975, 673 pages.

Abstract: This document (Volume 1, 2, and 3) presents a compilation of industry-wide methods in aerospace strength analysis that can be carried out by hand, that are general enough in scope to cover most structures encountered, and that are sophisticated enough to give accurate estimates of the actual strength expected. It provides analysis techniques for the elastic and inelastic stress ranges. It serves not only as a catalog of methods not usually available, but also as a reference source for the background of the methods themselves. Volume 1 (37.1 MB) Section A: Introduction, stress and strain, loads, combined stress, and interaction curves. A0.0.0 Downloadable hyperlink TOC, asm-A000.htm, 0.02 MB. A1.0.0 Stress and strain, asm-A100.pdf, 1.31 MB. A2.0.0 Loads, asm-A200.pdf, 0.35 MB. A3.0.0 Combined stresses, asm-A300.pdf, 0.98 MB. A4.0.0 Metric system, asm-A400.pdf, 1.03 MB. Section B: Methods of strength analysis. B1.0.0 Joints and fasteners, asm-B100.pdf, 2.14 MB. B2.0.0 Lugs and shear pins, asm-B200.pdf, 0.60 MB. B3.0.0 Springs, asm-B300.pdf, 1.08 MB. B4.0.0 Beams, beam tables, asm-B400.pdf, 1.32 MB. B4.5.0 Plastic bending, asm-B450.pdf, 1.27 MB. B4.5.5 Plastic bending curves: stainless steel, asm-B455.pdf, 6.03 MB. B4.5.6 Plastic bending curves: alloy steel, asm-B456.pdf, 2.82 MB. B4.5.7 Plastic bending curves: titanium, asm-B457.pdf, 0.96 MB. B4.5.8 Plastic bending curves: aluminum, asm-B458.pdf, 5.06 MB. B4.5.9 Plastic bending curves: magnesium, asm-B459.pdf, 1.31 MB. B4.6.0 Beams under axial load, asm-B460.pdf, 1.38 MB. B4.7.0 Lateral buckling of beams, asm-B470.pdf, 0.50 MB. B4.8.0 Shear beams, asm-B480.pdf, 1.76 MB. B5.0.0 Frames, frame tables, asm-B500.pdf, 2.47 MB. B6.0.0 Rings, asm-B600.pdf, 4.68 MB. Volume 2 (28.7 MB) Section B: Methods of strength analysis (cont.). B7.0.0 Thin shells, asm-B700.pdf, 0.74 MB. -- Slightly faded copy. B7.1.0 Membrane analysis of thin shells of revolution, asm-B710.pdf, 0.74 MB. B7.1.2 Dome membrane analysis, asm-B712.pdf, 1.18 MB. B7.2.0 Local loads on thin spherical shells, asm-B720.pdf, 1.70 MB. B7.2.2 Local loads on thin cylindrical shells, asm-B722.pdf, 0.86 MB. B7.3.0 Bending analysis of thin shells, asm-B730.pdf, 2.43 MB. -- Some tables faded! B7.3.4 Bending analysis of thin shells (cont.), asm-B734.pdf, 0.91 MB. B8.0.0 Torsion of solid sections, asm-B800.pdf, 0.85 MB. B8.3.0 Torsion of thin-walled closed sections, asm-B830.pdf, 0.73 MB. B8.4.0 Torsion of thin-walled open sections, asm-B840.pdf, 0.97 MB. B9.0.0 Plates, asm-B900.pdf, 1.72 MB. B9.4.0 Plates (cont.), asm-B940.pdf, 1.76 MB. B10.0 Holes and cutouts in plates, asm-B970.pdf, 0.73 MB. B10.2 Large holes and cutouts in plates, asm-B972.pdf, 0.96 MB. Section C: Structural stability analysis. C1.0.0 Long columns, short columns, crippling, asm-C100.pdf, 1.06 MB. C1.5.0 Torsional instability of columns, asm-C150.pdf, 1.18 MB. C2.0.0 Stability of flat plates, asm-C200.pdf, 1.20 MB. C2.2.0 Stability of curved plates, asm-C220.pdf, 2.59 MB. C3.0.0 Stability of shells, asm-C300.pdf, 0.52 MB. C3.1.0 Stability of cylinders, asm-C310.pdf, 1.96 MB. C3.2.0 Stability of conical shells, asm-C320.pdf, 0.80 MB. C3.3.0 Stability of doubly curved shells, asm-C330.pdf, 0.96 MB. C3.4.0 Computer programs in shell stability analysis, asm-C340.pdf, 0.87 MB. C4.0.0 Local instability of flat panels, asm-C400.pdf, 1.31 MB. Volume 3 (21.4 MB) Section D: Thermal stresses. D1.0.0 Thermal stresses, asm-D100.pdf, 0.54 MB. -- Some characters faded. D3.0.0 Thermal stresses in beams, asm-D300.pdf, 0.99 MB. -- Some characters slightly faded. D3.2.3 Thermal stresses in beams (cont.), asm-D323.pdf, 1.07 MB. D3.7.0 Thermal stresses in plates, asm-D370.pdf, 1.39 MB. D3.8.0 Thermal stresses in shells, asm-D380.pdf, 1.88 MB. D4.0.0 Thermoelastic stability, asm-D400.pdf, 1.20 MB. D5.0.0 Inelastic thermal effects, asm-D500.pdf, 0.55 MB. D6.0.0 Thermal shock, asm-D600.pdf, 2.21 MB. Section E: Fatigue and fracture mechanics. E1.0.0 Fatigue, asm-E100.pdf, 1.06 MB. E1.3.0 Fatigue (cont.), asm-E130.pdf, 2.43 MB. E1.5.0 Fatigue (cont.), asm-E150.pdf, 0.76? MB. E2.0.0 Fracture mechanics, asm-E200.pdf, 1.49 MB. E2.4.0 Fracture mechanics (cont.), asm-E240.pdf, 2.21 MB. Section F: Composites. F1.0.0 Composites, basic concepts, asm-F100.pdf, 0.77 MB. F1.2.0 Mechanics of laminated composites, asm-F120.pdf, 0.74 MB. F2.0.0 Strength of laminated composites, asm-F200.pdf, 0.24 MB. Section G: Rotating machinery. G1.0.0 Rotating disks, asm-G100.pdf, 0.55 MB. Section H: Statistics. H1.0.0 Statistical methods, introduction, asm-H100.pdf, 0.56 MB. H1.2.0 Statistical methods, measuring performance of a material, asm-H120.pdf, 0.66 MB.

Reply to
cavelamb himself

OK. Now I got you. The image were several rows of bolts (classical sheet overlapping joint) that are calculated more like rows of rivets. Not yet fully convinced that a stronger bolt has an disadvantage (except the price). *) But anyhow, the point is more the hole in the sheet and its projected surface area that has to accept a certain pressure (when shearing) and you can't make the bolt smaller because of the pressure. The clamping force is (almost) ignored in that setup. But maybe that was an old approach in design. I know, that joints like these are now glued (to take the shear) and screwed/riveted) to take the peeling forces.

*) But I also got your point about distributing load and what happens if one joint fails and you get an avalanche failure of the neighboring joints.

Nick

Reply to
Nick Mueller

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