Shear Stress Though Tapered Pipe Threads 2

[tc7]

I am trying to calculate the maximum pressure a threaded pipe fitting may withstand when the fitting has a NPT thread (1/4 -18 NPT). The failure mode I am worried about is excess shear load through the thread teeth. For this calculation I cannot determine shear area though the threads (because of the taper), nor can I easily determine length of engagement for an NPT thread.

Thanks to anyone who can provide advice.
tc7

 

[MintJulep]

If your pressures are really high enough that you need to worry about this you probably shouldn't be using tapered pipe threads in the first place.

 

[mrainey]

This free software will, among many other things, give you full dimensional data for NPT threads.

http://bellsouthpwp.net/r/s/rsnmar/METhreads.zip

 

[Cockroach]

This is a problem near and dear to my heart. You are correct in surmizing thread inclination as an issue, but for pipeline threads, there are other concerns.

This does not mean the problem is insurmountable. With good success, I have been applying Unified National specifications to NPTs, noting the tooth profiles are virtually identical, i.e. 60 degree vee type threads. The central problem is to "rectify" the threading inclination to an "equivalent UN profile", bastardized threads are allowed for computational purposes.

In your case, 1/4 NPT with an 18 TPI is virtually midway between 1/2 - 18 UNS-2G and 9/16 - 18 UNF-2G threading profiles. What I suggest is that a thread mid range between pin specifications behaves "somewhat like" that of a 1/4 NPT when loaded in shear.

Therefore it stands to reason that the 1/4 MNPT with major, pitch and minor maximum diameters of 0.536, 0.492, 0.447 inches respectively has the same shear area as a 17/32 UN thread, 18 TPI with thread classification tolerance near that of a coarse profile. I utilize a thread program, computations have been standardized or looked up in the Machinist Handbook, and obtain an equivalent major, pitch and minor pin diameter of 0.530, 0.494, 0.465 inches. What I am suggesting is that the 1/4 MNPT thread behaves like 17/32 - 18 UNC-2G.

Apply the threading computations for Unified National, ANSI B1.1-1982, noting that any thread pin with shear prior to the box (i.e. shear area of the screw is LESS than that of the nut) for maximum internal minor diameter of 0.484 inches, minimum external pitch diameter of 0.490 inches and find the screw shear area is 0.854 in^2 per unit length thread engagement.

When a NPT thread seals, although all threads are engaged between pin and box, typically only a few threads act to seal the joint. In the literature, computations use only one single thread. Therefore the effective length is around 0.047 inches, resulting in a screw shear area of 0.041 in^2. Naturally this is purely ficticous, when the threaded joint fails in shear we would expect all threads along the effective length of engagement to let go similtaneously. For the purpose of calculation, we recognize the single thread shear area method.

Hence, applying normal fracture mechanics, the pressure required to load a MNPT joint to failure is simply the product of pin shear area, material yield in shear divided by area of the thread seal, 0.492 inches in our case. The resulting mathematics suggests a pressure of 4700 psi assuming your material is Atlas T316 Stainless Steel. (TY = 22 ksi, Average Shear Theory)

Consulting industry specifications such as Swagelok, Parker-Hannifin, etc typical joint operating pressures are 4500 psi typical, never to exceed 5000 psi as a rule of thumb. Apparently our computation bears this out.

I hope this helps you out. My gut feeling with stainless fittings is to stay below 3000 psi.

Kenneth J Hueston, PEng
Principal
Sturni-Hueston Engineering Inc
Edmonton, Alberta Canada

 

[EdDanzer]

Check thread378-49291

 

[EnglishMuffin]

For what it's worth, I think MintJulep is right on. I think you should only use taper threads for low pressure on new designs these days. "O" ring fittings are the way to go. But once in a very great while, you might have an odd case where nothing else but a taper thread will fit.

 

[deanc]

Info. See ASME SecVIII Div1 UG-43(e) or SecI PG-39.5

 

[Fracball]

Why then, are the Parker and SwageLok books full of items with NPT threads that have a static operating pressure greater than 3,000 psi? Numerous fittings with NPT threads shown in these books have ratings in the 8,000 to 10,000 psi range. Based on these NPT thread ratings from fitting manufacturers, my company routinely uses fittings with 1/4 NPT threads in applications with a static operating pressure of 10,000 psi. Additonally, we have conducted tests on this particular thread size and found that its external threads failed at about 4 times the rating given by the manufacturer. For higher pressures (greater than 10,000 psi) the so-called "Autoclave" threads are a suitable option.

 

[Cockroach]

The issue of using a repeatable threaded joint with a NPT profile at an elevated pressure is with the increase in probablility of a wall crack, that is, stress induced failure originating at the root of the thread. It is very difficult to pull a threaded joint apart, agreed, but threads can also fail because the material below the minor diameter could not withstand the loading.

There are several outstanding reasons for it, the mechanics of the thread necessarily imply swaging the joint to specifications in order to provide a seal capable of carrying 10 or 20 ksi. Just gripping the pipe will necessitate external wall scaring while torsion is applied. Then there is wall allowance; obviously the joint must be able to withstand the stresses, i.e. Theory of Elasticity for Pressure Vessels. Typically the annulus are small, you need thick walls, perhaps the application cannot withstand choked or resticted flow, i.e. inlet feed to a pump. And what about shear? NPT threads cannot withstand transverse loading, they have a tendancy to bend at the last thread. The industry uses the notion of a fotbol to stabilize the joint should you need to carry weight lateral to the connection. And of course there is a material concern. Typically ferrous connections govern higher pressure ratings, stainless sometimes and brass never. Oilfield applications, my personal experience, do not use brass, yellow metal is forbidden since geological formations begin to yield sulfide bearing gases as the water tables begin to rise. Stainless is suspectable to pitting, chloride radial surface attack. I guess my judgement is one biased upon as a seasoned professional.

You are correct however, in noting prominent companies using elevated pressures. However, almost all note the maximum fitting working pressure is limited to the lowest NPT rating! Female pipe end ratings are typically much lower than their male counterparts of a given size, this is due to the geometry, inner/outer diameters for female threads are larger than those for male threads, i.e. higher stresses.

Kenneth J Hueston, PEng
Principal
Sturni-Hueston Engineering Inc
Edmonton, Alberta Canada

 

[Fracball]

Getting back to tc7's concern, how about using the NPT thread dimensions that correspond to the hand tight plane (or even the wrench tight plane) to calculate external tensile stress area, external shear area and internal shear area? From this, the applicable stresses (and thus strength) could be calculated. In other words, pick a plane along the length of the thread for calculating the needed tensile and shear areas. The Machinery's Handbook gives various length of engagement dimensions for NPT threads.

 

[tc7]

Frac-
I initially went down the path you are suggesting, but I could not find all the particulars on thread geometry for NPT's. The ASME B1.20.1 does not so thoroughly identify thread features as does say the H28 spec or B1.1 for UNC and UNF. So I was forced into some assumptions that I was not comfortable with for the calculation of shear area.

Cockroach -
I like the idea of your approach very much - it seems practical and no one can deny your experience with it, although I am not clear on some of your terminology. What do you mean by "pin" and "box"? More importantly, you stated the NPT thread was "mid range between pin specifications" of the 1/2-18 UNS-2G and 9/16-18 UNF-2G. Please elaborate on this. Also I am not clear on the "G" modifyer on your thread designations. I am only used to seeing "A" or "B" indicating external or internal thread.

Thanks again in advance.
tc7

 

[Cockroach]

a) Generic terms of pin, box are used in the oilfield to denote "screw" and "nut". When the term "bastard" is used, it means that the thread is not typically listed in a recognized specification or reference.

b) Recall that our methodology is to find an equivalent Unified National thread that is representative of your 1/4 NPT profile. Since the most important dimension of a thread is the pitch diameter, I suggest finding a UN thread closely matched by PD. The problem is that the pipe thread is inclined, therefore I use the handtight engagement as the "mean" pitch diameter for the NPT, 0.49163 inches in our situation. This is slightly larger than a 1/2 - 18 UNS-2G, but slightly smaller than a 9/16 - 18 UNF-2G thread. Suppose we split the difference and use the average between the two threads. This means I am searching for a thread with pitch diameter (0.463+0.525)/2 = 0.494 inches, i.e. mid range specification. This leads to a 17/32 bastardized thread which must have the same pitch as the NPT, 18 TPI in our case. Finally, pick the UN thread class according to NPT thread height, 0.04444 inches. The one I picked was 0.03250 inches, there may be one that is closer, but I ran with 17/32 - 18 UNC-2G.

c)Yeah, sorry about that! "G" references the thread fit, in our case class 2, which tells me something about the thread tolerancing applied in the calculations. You are absolutely correct in noting that "A" is the pin (screw-external) and "B" as the box (nut-internal). For UN threads we are permitted up to five (5) classes, 2 is what you would find as a commercial available size for example, i.e. off-the-shelf.

Just as a quick note, NPT profiles are specified by ANSI B1.20.3-1976, which is a revised designation of ANSI B2.2-1968. This same specification was reaffirmed in 1998. The layout is almost exact to that for UN threads, ANSI B1.1-1982 EXCEPT there is absolutely no appendix discussing fracture mechanics on thread performance. I find this curious, have been perplexed by it, inquired at several committee meetings to no avail. I remain perplexed and since, developed this method. You may find an old inquiry I floated in this website regarding this issue, but I have never seen, heard or otherwise given information on a solution to this problem. What is clear to me now is that others have these same problems!

Certainly, my pleasure. I hope this works out for you.

Kenneth J Hueston, PEng
Principal
Sturni-Hueston Engineering Inc
Edmonton, Alberta Canada

 

[chicopee]

I don't know if I am off base with your question, however I would not worry about shear stress thru the remaing wall thickness at the threaded joint. I would be more concerned about the hoop stress at the treaded joint.
Or are you refering about the shear stress resulting from the pipe and fitting joint under either tension or compression?

 

[EdDanzer]

Please clarify the ID of the pipe, material yield strength of both the internal and external thread, the OD of the external thread. Are you a worried about stripping the thread by pulling the 2 pieces apart, pressure rupturing the pipe or thread or is the pipe loaded in shear near the thread by another part?

 

[tc7]

Ed & Chic-
The point of my question was not in regard to hoop stresses or external tension/compression loads or loads applied transverse to the thread axis. Although these are valid concerns and I will address them in my design. My concern was with regard to the thread strength, i.e., due to shear loads through the threads which would be associated with pressue loads applied to a closed end fitting such as an instrumentation or a shut off valve.
tc7

 

[EdDanzer]

Unless the NPT thread has no hole and one of the threaded parts is of low yield, shearing the thread axially will be your least concern. To shear a thread axially takes about the same force as punching a hole the pitch diameter of the thread through a plate the engagement length in thickness with the yield strength of the weakest material.
As I stated in a prior thread we have pressure transducers and gauges that have operated at 20,000 psi and you can buy these off shelf. Yes they need to be used with 60,000 psi minimum yield materials. In the testing we have done you will typically burst the tube before the thread will fail.