Pin bearing in greatly oversized hole in plate 1

[mad13]

I have a unique steel pin connection that I am unsure of how to analyze. I have a 1" diameter pin connected to a hasp which is loaded by a 5 ton hoist. This pin passes through a 2-1/2" diameter hole in a 1/2" thick plate. The plate is welded to a hoist beam. I am tasked with load rating the hoist beam and the connected plate. My problem is that AISC section D5 limits the application of that section to pin holes no greater than 1/32" larger than the pin diameter. In addition, my "a" dimension (shortest distance from hole edge to outside material edge) is 1.75", less than the 1.33 x b(eff) = 2.168". Therefore I believe I can not use the criteria in this section to check for tensile or shear ruputure. The code gives no direction of what section of the code applies failing these two dimensional checks.

If I apply section J4, because of the wide width of the plate, it seems that only block shear would control, due to the relatively short edge distance in the direction of the force. My inclination in that case would be to use the "a" dimension times the plate thickness for A(nv) and A(gv), with no A(nt), because there would only be 1, not 2 shear planes, because the smaller diameter pin causes a point load on the hole wall rather than trying to fail a section of the plate of the width of the hole in the plate. I would attach a picture, but the software for uploading files in this forum does not seem compatible with this computer's security settings.

Can anybody help?

 

[ToadJones]

I didn't even read the whole post....but...

You may need to check this for "below the hook lifting devices"...not sure of the code but I beleive it is ASME

 

[JStephen]

There is an ASME specification, which is fairly general, and then there is ASME BTH, which is a design guide for below-the-hook equipment designed to that ASME specification. The specification may or may not apply to your situation; however, the BTH design guide does have equations for allowable loads of pins in holes without the 1/32" limitation (and I don't recall if there is ANY limitation right offhand).

One thing that is not covered by the BTH, is that with a hasp (or shackle?) that is much wider than the plate is thick, it could hang at an angle, and the design assumes the pin is perpendicular to the plate.

 

[Ron]

For analysis purposes, if you draw a 1" diameter circle tangent to a 2-1/2" diameter circle, you'll see that the shear will be concentrated along a line, rather than spread into 2 lines, so essentially whatever your shear load is, you'll have to consider the whole load, not half. Further, because the plate is relatively thin, you might get hole distortion (out of plane) because the hole is so large relative to the pin.

Lastly, your contact pressure on the small pin with large diameter hole will be high, thus limiting your load capacity.

Why can you not fill the 2-1/2" dia. hole and re-drill for the 1" pin?

 

[dicksewerrat]

Cut the old plate off. Weld a new plate with proper size hole. Then do your calcs with the smaller hole to get load rating. This sounds bad sloppy lifting device. OSHA may not like that thing the way it is now.

Richard A. Cornelius, P.E.

http://WWW.amlinereast.com

 

[SAIL3]

looks dicey to me, the herzian stress will be off the charts......attached is one of many articles on the subject...you choose the allow max stress..

 

[dhengr]

A good sketch with some dimensions and loads, etc. would really be helpful, if you want any meaningful discussion on your problem. I’ve seen that general type of problem dozens of times. but I don’t have the latest Eds. of AISC 360, so now, all I have to do is conjure up what “a,” b(eff), A(nv), A(gv), no A(nt), one or two shear planes, and a hasp mean with respect to this particular design problem, and we could probably have a discussion on your problem. While much of the theory and thought process embodied in AISC 360 may generally be applicable, I doubt that their cookbook code approach and formulas will be. You really have to read btwn. the lines to understand why they do some of what they do and how that sets some of the limits they set. OSHA and ASME have some requirements and specs. on these kinds of things, as Toad suggests. Do you know what kinds of materials the various parts are made from; Fy, Fu, E, etc., what about weld sizes and strengths?

You’ve got some of the right terminology, apparently from the latest AISC 360 specs., but unfortunately they do not cover every conceivable detail that can be dreamed up. In which case you have to fall back on Strength of Materials, Theory of Elasticity, and good engineering judgement and experience which they haven’t quite figured out how to codify yet.

I don’t understand what you mean by a hasp. Is this the base of the hoist, with two plates, one fitting on each side of your .5" pl. and then through pinned to it? The 1" dia. pin bearing on the 2.5" dia. hole in the .5" pl. is a Hertz bearing stress problem. Immediately under the pin you will have yielding in bearing and high shear stresses. There can also be some tensile stresses and beam bending stresses depending upon the proportions of the various components in the detail. You can get rupture on a single shear plane which is parallel to the length of the pin and at the pin center. This is due to the bearing, shear and bending normal (+ or -) stresses immediately under the pin, or it at least originates there. And, you can get what I think you called your double shear planes which are essentially the pin dia. apart, and otherwise basically parallel to the pin length again, but probably not on what you are describing. This problem is much more akin to the analysis of an eyebar than it is to what I think you are looking at in AISC, but the large difference in the pin dia. and the 2.5" hole complicates this problem. Putting the 1" pin inside a piece of heavy walled mech. tubing will change the relative diameters which are so critical to this problem; then the pin and mech tubing are put in the 2.5" dia. pl. hole, in effect a bushing on the pin. But, again, I can’t see your detail from here.

 

[steele6707]

Attached is a page from ASME BTH-1-2008. I believe I paid a little over $50 (US)and purchased from ASME. This is somewhat similar to AISC 9th Edition and provides for additional adjustment factors to allowable stresses based on the usage of the device.

 

[BAretired]

Using Article 13.10 of CAN/CSA 16-01, for Fy = 300MPa, Br = 4215N, 4.2kN or 948# where Br is the factored bearing resistance.

I like the idea of using a bushing around the pin as suggested by dhengr.

For a 2" dia. bushing in a 2.5" hole, Br = 25.3 kN or 5,690#.
For a 2.25" dia. bushing, Br is increased to 56.9kN or 12,800#.

The expression used in CSA 16-01 is for bearing of expansion rollers or rockers and is as follows:

Br = 0.00026 phi*{R1/(1-R1/R2)}*L*Fy^2 (in Newtons)

phi is taken as 0.67 in the Canadian Code.






BA

 

[saplanti]

In case the pin is going to work for large number of applications, and the pin rotation may be exist during the lifting/force application, the hertzian stresses are very essential factor to consider. The only way to reduce the hertzian stress is the make the pin and lug diameter closer. The closer is the better.
Therefore I suggest you NOT to ignore the clearance requirements, if you involve with applications above.

In some static applications (pin does not rotate) the bearing area allowable with the area (the pin diameter * thickness of the lug) is acceptable. When the force apply on the pin, the hertzian stresses gets a lot larger and yield starts. The yield continues under the load until the required bearing area occurs under pin on the lug. In case you still have sufficient strength on the lug against shear and bending stress the system works. However there is no warranty on this type applications, and codes are pushing to use a proper pin lug connections by adding a maximum tolerance on the clearance.

Check different standards for the pin connections, and understand the the requirements, where and why they come from. Why a maximum dimension to lug thickness ratio is asked to use.

SAIL3 has provided a good guide for the Hertzian stress calculation to make you sleep good at nights.

But still you are obliged to use the code rules for the connections.

Cheers,

Ibrahim Demir

 

[mad13]

Attached is a drawing and picture of the specific situation. I do have the ASME BTH - 2008 specification that ToadJones and steele6707 referred to above. The only thing that this does not address is the "hertzian" contact stresses referenced above by dhengr and saplanti. The customer is not interested in changing the connection (at least at this time) - they just want to know the maximum load this hoist beam can safely lift in its current state. If that value is not acceptable to them, then they will ask me to supply alternatives to improve the capacity - such as changing the relative diameters of the pin and hole.

 

[dhengr]

Mad13:
I think Ozgurx meant to say ‘it will yield too much and therefore it’s not a good design,’ instead of ‘it won’t yield a good design.’

If the pin is 1" dia., then that’s most likely a 7/8" shackle and it has a working cap’y. of about 6.5 tons and a FoS of 4-5. But, you have to check the pin in your case because it is really not intended to be used with a .5" pl., when its clear span btwn. the shackle legs is about 1 & 7/16", although they often are misused this way. Another thing you should check is the whole condition of the system, it really looks pretty well weathered. When that .5" reaction pl. gets rusted or abused around/in the 2.5" holes or on the sides of the pl. adjacent to the pin contact points, this becomes a potential stress raiser in a very highly stressed area on the pl. Regular inspection and care of the pl. hole edges may be as important as anything because any stress raisers on these edges can be crack starters.

It appears to me that your problem is primarily a Hertz bearing stress problem under the 1" dia. pin in the 2.5"dia. hole in the pl. Roark’s, Formulas for Stress and Strain has some fairly good material on this topic, although it’s hardly the final word on the subject. It’s a combined stress problem immediately under the pin which can lead to ripping or fracture. And, it is going to end up being an engineering judgement call, because none of the codes really cover this. They all kinda dance around this issue and normal address it by limiting the dia. difference btwn. the pin and the hole; you say AISC says this dia. difference can’t be more than 1/32", but that may even be too much in smaller pins. Some of their limitations address several potential problem areas at once. And, that’s what make some of their formulas so difficult to decipher. Secondly, that loose fit shackle on the .5" pl. can lead to out-of-plane loading (in effect a twisting load) and buckling of the pl. At a min. I would add some washers (maybe 7/16" thk. x 1.06" o.d. x 4" i.d.) on the pin on either side of your .5" pl. to keep the shackle centered on the pl. and prevent it from tipping the pin off horiz. in the hole. My guess is that the failure mechanism in your case would be single plane fracture right under the pin, that’s ASME’s formula (3-49) posted by Steele6707. Look at ASME’s pl. bucking (dishing) criteria also. I certainly would be interest in seeing a few pages on either side of pg. 25, so I could see all the variable definitions and commentary on the subject. Could you post them or this section of the code, so I can have a look? I didn’t have the advantage of the ASME BTH-1when I first started doing these problems, and I don’t have this Ed. of it now.

ASME BTH-1 is the better code to follow on this problem. Good luck.

 

[mad13]

I have analzyed the plate per section 3-3.3 of ASME BTH - 2008, and obtained an allowable load value of 19.2 kips, controlled by single plane fracture strength. I have also attempted to apply the hertzian contact stress criteria, summarized in the article forwarded by SAIL3, and only obtained an allowable load of 430 lbf (unless I am incorrectly applying this formula). As far as the pin is concerned - I am not tasked with analyzing the pin, hasp or anything other than the lift beam and plate - altough I will pass along the recommendation to either increase the pin diameter to match the hole diameter, or supply washers as dhengr has suggested.

I have attached my analysis, along with the whole ASME-BTH reference as requested by dhengr.

 

[saplanti]

This is another discussion on the issue, and may be useful for you.

http://www.eng-tips.com/viewthread.cfm?qid=273125

Regards,

Ibrahim Demir

 

[SAIL3]

some good points have been addressed...
not surprised that the hertzian stress governs....as saplanti mentioned the material will yield resulting in incresed contact area of the pin until equilibrium or failure occurs...some yielding I can live with but don't know the safe cut-off point.
As commonly happens in rigging...they have a bin with different size shackles and just reach in and grab whatever one will fit..I try and force them to use the correct shackle by a combination of lift lug pl thickeness and pin hole size that will fit a certain standard shackle size....this pl thickeness can be increased by adding(welding)) pl washers(doughnuts) on each side of the lug, if neccessary ...this way I don't have to worry about unintentional bending in the pin...

 

[saplanti]

SAIL3,

You still should worry about unintensional load transfer with shackles. In accordance with all the standards/rules you need to make sure the shackle pin centralized during the lifting. For that you may require to put additional washers inside the shackle area on the pin.

It is your responsibility that you record this on the drawings if the internal size of the shackle against the lug thickness at the connection leave large gap which may cause the load concentrated on one side.

I additionally have a long discussion with Australian Standards on the reducing the tolerances of the shackle pins as well to maintain the reasonable and controlable gap. But they said they are mainly locked by the American manufacturers, so they can not do anything about it. If the engineers specify them by ignoring the pin tolerances they are making big mistake for the pin calculations. In the Hertzian stress check you need to consider the worst diameter which is always the minimum diameter into consideration. But I do not know what the standard you are using for the job is saying about the maximum gap rules at the moment with the shackle applications on lugs.

Hope it helps,

Ibrahim Demir

 

[saplanti]

Please ignore the part of second last sentence "..into consideration". Sometime I mess up with the sentences and do not see them before launching.

Ibrahim Demir

 

[SAIL3]

The final design of the lug should be checked/matched with a standard size shackle ie:
Say pin size is 1 1/4" dia....this would correspond to a 1 1/8 size shackle which has a clear inside width of 1 13/16"...
now if the lug pl thickeness is , say, 3/4" , this would leave a gap of 1 1/16" between the lug and shackle legs causing bending or off-centered load on pin.....
solution..either increase the thickeness of lug pl or weld on pl washers(doughnuts) to lug..
the solution mentioned by saplanti of adding loose washers on the inside would certainly center the load but would not address the bending in the pin..also in my experience requiring loose pieces to be used does not guarantee that they will actually do so...
I would live with a 1/8" gap or maybe more depending on the situation..
also a 1/8" oversize hole in the lug sounds reasonable to me...

 

[SAIL3]

meant to say 1/16" oversize hole in lug would be reasonable....

 

[Teguci]

I am on board with what SAIL3 is saying. It is good practice to size the lifting lug to exclude the wrong size shackles. None of my lifts have failed, but I have talked with people who were around when a lift went bad quickly. Don't skimp on the welds either.

ASME BTH 3-3.3.5 - Calls for a hole size no greater than 110% of the pin diameter. Failing that criteria, BTH requires further analysis regarding the extra clearance.