DNV Pad Eye check. 1

[truckdesigner]

Something I've pondered for a long time:

Why don't the DNV rules (2.7-1) check for "tensile tear-out" and only "shear tear-out"? As I understand there are a number of failure modes of a lug, bolted connection etc. Typically the 3 checked are shear, tensile and bearing.

The allowable stress for shear is less than for tensile, but if the hole centre is far enough back from the lug edge the calculated failure mode does become tensile tear out - and not shear tear-out.

A good example of this is demonstrated at this website:

http://expeditionworkshed.org/workshed/bolted-steel-connection-failure-plate-tearing/

Also - see attached.


Regards.

 

[JAE]

I have no idea what DNV is (might help to define it).

Your link and your attachment don't work for me either so I can't help you based on that.

It seems that typically shear is the main check for single bolts or pins. Tension would require a section perpendicular to the force direction and usually for single bolts and pad eyes that just doesn't control based on usual geometry of the plate.

 

[ztengguy]

Dept of Notor Vehicles?

 

[JAE]

Ha!

 

[dik]

It's a Norwegian standards group that has some excellent technical publications (generally free).

Dik

 

[NOLAENG]

Why not ask DNV? It seems like they would be the only people who can answer this. All of the lift eye/pad eye/lifting lug design standards I have used check the three you mentioned. AISC, ASME, internal company standards, etc.

 

[JStephen]

"Why don't the DNV rules (2.7-1) check..."

Some speculation on the original question:
Perhaps because that mode of failure is considered a general tensile failure, applicable to any member in tension, and not specifically padeyes. For example, ASME BTH includes allowable gross and net tension of tensile members, but then has a separate section on pin connections which does not include this mode of failure, either.
Perhaps the standard writers just didn't think of it.
Perhaps in practical applications with the padeyes in question, that mode of failure never governs.
Perhaps the standard is intended as a minimum requirement, but is not intended as a complete cookbook of every possible design aspect that would ever need to be checked.

You might consider submitting proposed rule changes or checks on the issue to the standard writers for future revision if needed.

 

[EngineeringEric]

Was this question posted just an excuse to show off some fun clips of materials failing. Thank you for the link! i lost a good 30mins of billable time watching some the videos there

I reiterate above, ask the DNV what they are looking for.

 

[dhengr]

Truckdesigner:
I suspect that the little bit you attached from the DNV std., is their prescriptive method for that shape of pad eye, and they do say that your can (should when necessary) refine your design if you wish, with more detailed calcs. and methods. They do talk about a load factor of 3, and that concentrated stresses should not exceed 2Fy. For that shape, you can often see by inspection, if a tension failure is likely. The stressed area is often more than double the steel area for the single plane of shear failure and the yield stress in tension is Fy while the shear yield stress is only about .58Fy. Your videos show several of the failure modes quite well. You can approach a tension failure when the two side legs (steel areas) get much smaller in radius vs. pin hole dia., as in some eye bars. Please verify this with someone on the DNV committee or further review of their std., but my thoughts are as follows: Their first equation is a tension failure calc., on the horiz. center axis of the pin hole, with the denominator being the net area on that plane. And, with a load factor of 3 they ignore some of the small stuff. With a pin which doesn’t fill the hole, there can be a secondary bending moment tension component on the pin pl. at the hole edge, on that same plane, like the moment on a lifting hook. The equation for the single plane shear failure in line with the load would be something like, .58Fy ≥(3*RSL)/[(H - D/2)(t)]. But, here again, they leave out some of the small stuff, since on this same plane and at the edge of the hole, you will have max. bearing stresses from the pin, and thus very high tri-axial stresses where this shear failure can start to unzip. This mode of failure just starts to unzip and open up on that single plane. That is also the area on the pin hole which takes the most abuse, nicking, yielding, deformation, particularly from smaller dia. shackle pins, so be careful. Their second formula comes from the Hertz bearing stress problem and is highly dependant on any small difference in dia. btwn. the pin and hole, as they note, no more than 6% difference. There are two more primary failure modes, and those are a double shear plane failure mode where a plug about 2/3rds the pin dia. is pushed out, and the rest of the pl. spreads a bit to accommodate this. This is much dependant upon the shape of the pad eye pl. causing a min. shear area along with high secondary tensile stresses. Your shear failure video kinda shows this failure mode. It also shows the other failure mode, and that is dishing or buckling out of plane because the pad eye pl. is too thin. You can see the plate dishing, deforming out of plane above the bolt during failure.

Look at ASME BTH-1 “Design of Below-the-Hook Lifting Devices,” it goes through this same problem, with a slightly different tack on the problem. It seems to me that DNV has had a more detailed approach in some material of theirs that I’ve seen, although I don’t have a complete copy of their standard. They both end up with about the same results, and come from the same origins as far as Engineering Mechanics and Strength of Materials goes. Things like eye bars, lifting hooks, cylindrical shapes with various loadings, bolts & bolt holes near edges, all had considerable history and closed form long hand solutions, testing and study. The bearing stresses btwn. the pad eye pl. and the shackle/shear pin have their origin in the Hertz contact stress (bearing stress) problem, and are very dependant upon the relative dia. of the pin hole and the pin itself. Dig out some good Strength of Materials, Theory of Elasticity, Machine Design, etc. text books and you can wade through where this stuff comes from, and what the likely failure modes are.

As much as anything good quality detail and welding without any stress raisers and minimized stress concentrations make for good lifting equipment design. I always have them grind or machine a .125" +/- radius or chamfer on the corners of the plate, both in the hole and on the outer edges of the pad eye plate. This reduces the likelihood that these corners will get nicked and cause stress raisers. Have nice clean load paths into the structure the pad eye is fixed to. The pad eye pl. should not cross a web pl. below, without special attention to detail. Be very careful about welding around the two edges of the pad eye. This tends to cause notches at the four corners of the edges.