Acme Thread Stress - Drill String Connection

Probably not too concerned since the drill rod is usually tightened up to a shoulder on the bit or on the coupling.

Ted

The Acme thread form is perhaps the strongest connection geometry for our industry. A lot of bastardization techniques that I have used make use of a high torque shoulder, as suggested by HydTools, that limit bending since bearing load is transferred to that area upon make-up. I suggest that your combined stress is therefore, so high.

In the drill string, axial load is usually the least concerning, torsion the greatest. For example, in the load application of a drilling driveshaft, 85% of stress is attributed to torson, 12% to bending and the balance to axial load. This has been my experience with a major energy service company providing drilling solutions to the industry. Several FEA model support the API specifications that HydTools speaks of, say API 7, 7G, 5CT for alternate tubulars and the various thread forms thereof.

The 4 TPI, 3/4 TPF Acme thread is a personal favorite, never let me down yet. Highly robust, easy to cut, easy to QC/QA. Hope this works out for you.

Regards,
Cockroach

Thanks for the information Cockroach and Hydtools, however, I am still a little...confused...worried. Let me explain the situation more thouroughly. Due to the double shoulder connections (Vam, Grant Prideco, Etc)that are becoming more prevalent, it is becoming harder to fit parts into a Kelly Cock while maintaining the desired drift diameter. In order to maintain the drift diameters, we are using two piece bodies. The body connection in that past has been an API connection ( with non-standard bevel diameters) that are rated slightly higher in torsion and tension than the double shouldered connections. However, in this case, I am unable to use an API connection that is stronger due to O.D constraints. The tensile load rating and torsional load rating on the double shoulder connection are 1.2e6 lb and 68000 ft-lb respectively. I want to ensure that my body connection (4 tpi Acme) is 10% stronger. I guess the questions ( based upon these loads) are:

1. Do I simply need to check thread shear, tensile stress in the box and pin, bearing load on the faces that shoulder out, and torsional stress in the box and pin?
2. Or should I be checking bearing stress and bending stress in the threads and combining the appropriate stresses to obtain the Von mises stress at the thread root?

Any insight would be appreciated.

Heliakon

I would follow the API Specification 7 & 7G for the connection strength calculations and compare the same calculation to other geometries. In that way you could tabulate and cross compare your results to support your claim on performance.

Get both those specifications and write the mathematical model so you can compute each threading geometry. The equations are fully covered in that standard. It's really not a big calculation and definitely nothing to fret over! But you need to do the work.

Regards,
Cockroach

Thanks Cockroach. I will do that.

Cheers
Heliakon

Cockroach:
I’ve seen some of your calcs. in the past and you usually lay things out very nicely, cleanly and clearly. I’ve run across a few API specs. in my past life, but basically don’t know one from the other unless I’ve got them in hand. And, I agree with your statements about Acme thread strength and you general magnitudes of torsion, bending and tension stress in this joint. But, maybe a non-oilman’s take on coupling stresses and loads would help Heliakon’s understanding.

Once this joint is made up, it essentially acts as a continuous shaft, except for minor secondary loadings and stresses in the joint. The sum of the shear stresses/loadings in the threads equals the bearing stress/loads on the shoulders, plus or minus. These stresses and loads are quite high, as the material grade will allow, and essential make the shaft a continuous piece. For bending or tension to become significant issues they would have to overcome this high preload in making up the joint. Tensile loading is well distributed around the joint, but starts to counteract this preload, and in the extreme would start to open up the shoulder joint. Bending stress will vary across the shaft/pipe and will be additive or subtractive from the shoulder bearing stress or thread shear stress, depending upon which you are looking at or their radial location on the shaft. But, until these combined bending and tension stress start to overcome some significant percentage of the preload the joint remains tight (a continuous shaft) and bending or tension stresses are secondary stresses, but not to be ignored, either. Probably, experience and testing allows you or API to put numbers on what I called ‘significant percentage.’ In the structural business I would liken this problem to preloaded bolts in a tension or moment carrying structural joint.

Please flesh my thought out for Heliakon, as you see fit; or tell him I’m all wet and should mind my own business.

Thanks dhengr. I have found some reference material (Roark's Formulas) that has cleared up my issue. According to the 7th edition of Roark's, the transverse shear (due to make up torque) is the dominant force in the threads, due to the fact that the thread is essentially an extremely short beam with a great depth. This has cleared up what I thought was an overlooked stress in the thread itself.

Now dealing with this threaded connection makes sense.
Thanks all for your input.

Cheers
Heliakon