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In article , Gary Coffman says...

(NPT threads)

They're *tapered*. That means you don't have full depth threads
engaging full depth threads at any point in the joint.

I certainly agree with your conclusion, but NPT threads
are designed to have full thread engagement. The
taper angles match. Imagine a morse taper with the
ID and OD threaded.

Machinery's
Handbook says that only NPS threads should be used for axial
loading.

Strong agree on that.

Jim

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"jim rozen" wrote in message
...
In article , Ed Huntress
says...

You probably could approximate the strength if you calculated the
strength
of three straight threads, more or less, of the same size in the same
material.

Why three? The best way would be to use the smallest
wall section.

It's stronger than that, Jim. Three is a guess, but that's the range over
which there is a roughly balanced wall section on both sides of the joint.
Where the wall section is thinnest is at the ends; neither end is the
determinant of strength.

And the sharp threads do indeed reduce
the strength.

Oh, I wouldn't dispute that. It's just that it isn't necessary to show that
the joint is weaker than one with parallel threads. And it's always
problematic because it depends on the threading tool that cuts the thread.
The standard is one thing; actual strength is another, in other words.

In fact, I just took a look, and there are ANSI, ISO, and JIS standards for
mechanical strength of tapered threads -- apparently. I say "apparently"
because they're mentioned in the abstracts, but it costs money to get the
full standards, and I'm not buying. g

It should be noted that parallel pipe thread standards say "mechanical
joints," and tapered pipe threads say "sealing joints," but both mention
mechanical strength.

Ed Huntress

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In article ,
Ed Huntress wrote:

It isn't the sharp threads, Ned. It's the fact that the walls are weak at
each end of the taper, because one side or the other is thin at either end.
Only a couple of threads in the middle of the joint can produce the full
strength you would get with straight threads, and that isn't enough to
exploit the strength of the material. There aren't enough threads where the
material on *both* elements is equally thick.

Uh, in that respect the taper is actually an *advantage*, not a
disadvantage. The thin walls at the end of the taper, where there is
less strength, are also more stretchy, so they don't attempt to bear as
much load. The normal rule of thumb that only about three threads bear
the stress (and that it is actually concentrated on the first or last
thread) can be improved by using a taper.

(Consider, as a thought experiment, two tubes of the same OD and ID
screwed together using an extremely fine tapered thread, with the taper
extending all the way from the OD to the ID, so that looking at the joint
as a whole, there are no stress risers on the inside or outside. The
whole thing stretches like a solid mass; there is no stress concentration
on any first three threads or last three threads. Of course this
situation -- both OD and ID matching, and ultrafine threads -- is not the
situation we're considering here, and any other situation will be worse,
but it does serve as the extreme example of the advantages a taper can
bring.)


--
Norman Yarvin http://yarchive.net

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OK Stu , my profound apologies for being one of the reprobates ragging you about multiple posts.
Your project looks like a very creative job.

My curious mind forces me ask:

A: I don't recall seeing you mention why you built this rig Stu. Is there a handicapped person
living there who can't do stairs?

B: What are the alternate means for this person (if they are disabled) to safely get down and out if
there's a fire and the electric power is lost?

C: The offset position of that lifting pipe on the lift platform makes for quite a lot of bending
load on the joint in question. And maybe a bit of "bounce" when it stops at the top or when someone
steps on at that location. How is that pipe flange under the platform fastened to it so that it
won't tear loose.

D: I didn't "get" the safety brake concept from the photo, but it sounds interesting. Does it work
something like the jam washer on a screen door closer? Can you amplify the description for me? I
realize there's very little chance of it ever being needed, but did you actually test its
performance with a dummy load by snipping a temporary link in the winch cable while the elevator was
in motion?

Finally, I myself wouldn't trust the 3/4" black iron pipe threads for this job; not with a person's
safety depending on them. Assuming you can accomode assembling things in place with the platform
fastened to the pipe, I'd go with something better than hardware store pipe there and spend the
bucks to have it welded into a 1/4" thick steel plate (about one half the area of your platform) by
a certified welder, with a collar or sleeve welded to it and the plate as well.

Just my .02.

Jeff

Jeff Wisnia (W1BSV + Brass Rat '57 EE)

"If you can keep smiling when things go wrong, you've thought of someone to place the blame on."




Stu wrote:

Group,

First let me apologize for posting my question more than one time.
The problem was the Google was 'hiding' my posts and I repeated them
because I believed that they were not posted. I asked Google to
'splain.

Why did I ask this questiion?

See my homebuilt elevator that 'hangs' on the threads of a single 3/4"
black iron pipe.

http://www.imagestation.com/album/?id=4288894713 Homebuilt Elevator

Thanks again,

BoyntonStu

Jeff Wisnia wrote in message ...
Ed Huntress wrote:

"Stu" wrote in message
om...
Strength of 3/4" iron pipe threads.

Thanks,

Boyntonstu

Here ya' go, Stu:

3/4" iron pipe threads produce 32,450 ksi inverted shear strength. That is,
assuming you've torqued the joint to a value between 86.5 and 86.6 ft-lb,
backed off a half-turn, and re-tightened by 11/16 of a turn.

Enjoy!

Ed Huntress

You can get at least 23.76% more than that if you use unobtanium pipe.

Jeff




--

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Stu wrote:

What are you talking about?

Stu


I think he tried to copy a post response to you and got a message back from your "Automatic e-mail
verifier" Stu, the one for "Stu or Jan". It looks a somewhat like it might be of e-mail harvesting
scam. This is what it just sent to me:

************************************************** **********************************

Subject:
Please verify your Email address for us this one time. (Free) - 564018913810
Date:
Tue, 14 Oct 2003 12:49:02 -0400
From:
Automatic Email Verifyer (Stu or Jan)
To:
"Jeff Wisnia"

This is an automatic message from Stu or Jan.

In order to avoid SPAM Emails, We require that
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--
Jeff Wisnia (W1BSV + Brass Rat '57 EE)

"If you can keep smiling when things go wrong, you've thought of someone to place the blame on."





FuhhKyu wrote in message . ..
And he wants us to register so we can watch 15 minutes of his dirty
movies. I don't want any spam so I'm not gonna register.

Nor am I gonna register my gun


On Sat, 11 Oct 2003 04:24:33 GMT, Jeff Wisnia
wrote:



Stu wrote:

Strength of 3/4" iron pipe threads.

Thanks,

Boyntonstu

You just don't give up, do you Stu?

Jeff

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Jeff Wisnia

I use Emailbouncer.com to avoid and eliminate all SPAM, Porn, and
Virus attacks.

The first time Emailer must verify the email address one time. Else,
a SPAMMer using any email address could get to my Inbox.

If you sent me a valid Email without Emailbouncer I would know your
email address.

BoyntonStu





wrote in message ...
Stu wrote:

What are you talking about?

Stu


I think he tried to copy a post response to you and got a message back from your "Automatic e-mail
verifier" Stu, the one for "Stu or Jan". It looks a somewhat like it might be of e-mail harvesting
scam. This is what it just sent to me:

************************************************** **********************************

Subject:
Please verify your Email address for us this one time. (Free) - 564018913810
Date:
Tue, 14 Oct 2003 12:49:02 -0400
From:
Automatic Email Verifyer (Stu or Jan)
To:
"Jeff Wisnia"

This is an automatic message from Stu or Jan.

In order to avoid SPAM Emails, We require that
every person do a one time verification of
their Email address. This step should only take
a few seconds. Loading the page will tell you
that your Email address has been verified.
To save time, you do NOT have to load the entire page.

Simply click on the hyperlink below to deliver the Email
that you sent to me with subject :

Machinest Handbook lookup request (Pleeeeze)


Once verified, all future Email will immediately be delivered.

Thanks,

The Email Verification Autoresponder


http://www.email-bouncer.com/verify.cfm?564018913810

This email is to verify :
Key : 564018913810

************************************************** *****************

Jeff

--
Jeff Wisnia (W1BSV + Brass Rat '57 EE)

"If you can keep smiling when things go wrong, you've thought of someone to place the blame on."





FuhhKyu wrote in message . ..
And he wants us to register so we can watch 15 minutes of his dirty
movies. I don't want any spam so I'm not gonna register.

Nor am I gonna register my gun


On Sat, 11 Oct 2003 04:24:33 GMT, Jeff Wisnia
wrote:



Stu wrote:

Strength of 3/4" iron pipe threads.

Thanks,

Boyntonstu

You just don't give up, do you Stu?

Jeff

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In article 1JKib.12573$Eh3.6300408
@news4.srv.hcvlny.cv.net, says...
"Ned Simmons" wrote in message
...


Oh, come on, Ed. What's so different about pipe threads
that they don't "exploit the strength of the material"? The
sharp thread form likely causes some stress concentrations,
but other than that, I don't see much difference.

It isn't the sharp threads, Ned. It's the fact that the walls are weak at
each end of the taper, because one side or the other is thin at either end.
Only a couple of threads in the middle of the joint can produce the full
strength you would get with straight threads, and that isn't enough to
exploit the strength of the material. There aren't enough threads where the
material on *both* elements is equally thick.

If you look at an actual pipe fitting, unless it's one of
the thin "merchant" couplings, the wall is much thicker
than that of the pipe. I still maintain that the weakest
point is the pipe wall where it exits the fitting, per my
exchange with Jim.


You probably could approximate the strength if you calculated the strength
of three straight threads, more or less, of the same size in the same
material.


Here the NPT thread has an advantage over a UN thread. The
shear area is a function of the minor diameter of the
external thread, and for a tapered thread the minor dia of
the ext and internal thread are equal. Measuring in the
middle of the thread engagement on a 3/4 fitting, I get a
minor dia of about .92

(.92 in X pi) x (3 threads / 14 threads/in) = .62 in^2

Assuming a shear strength about 1/3 of UTS (very
conservative, I think)...

..62 in^2 x 20000 lb/in^2 = 12400 lb.

Which is pretty close to the yield of the full (unthreaded)
wall of the pipe.

Ned Simmons

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In article ,
says...
On Mon, 13 Oct 2003 20:22:56 -0400, Ned Simmons wrote:
Oh, come on, Ed. What's so different about pipe threads
that they don't "exploit the strength of the material"? The
sharp thread form likely causes some stress concentrations,
but other than that, I don't see much difference.

They're *tapered*. That means you don't have full depth threads
engaging full depth threads at any point in the joint. Machinery's
Handbook says that only NPS threads should be used for axial
loading.


The major, minor, and pitch diameters all follow the same
taper, so the thread form remains constant throughout the
length of engagement.

You must have a different MH than I do - mine (22nd ed)
says,

"While external and internal taper pipe threads are
recommended for pipe joints in PRACTICALLY EVERY SERVICE,
there are mechanical joints where straight pipe threads are
used to advantage."

And from Kent's Mechanical Engineer's Handbook (11th ed),

"Taper male and female threads are recommended for threaded
joints for ANY SERVICE. ... Straight male threads are
applicable only to special purposes, as long screws and
tank nipples."

My emphasis added in both cases.

Ned Simmons

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In article , Ned Simmons
says...

.62 in^2 x 20000 lb/in^2 = 12400 lb.

Which is pretty close to the yield of the full (unthreaded)
wall of the pipe.

Again, the trouble arises with the sharp V form
at the root of the thread, and also the die-cut
threads. Those are the two *worst* situations
to have when one is trying to optimize the
strength of the joint.

Each of those will substantially reduce the
load carrying ability of a connection like that.

A proper fastener will have a) rolled threads,
with b) a bit of a radius or flat at the root.

Jim

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In article ,
says...
In article , Ned Simmons
says...

.62 in^2 x 20000 lb/in^2 = 12400 lb.

Which is pretty close to the yield of the full (unthreaded)
wall of the pipe.

Again, the trouble arises with the sharp V form
at the root of the thread, and also the die-cut
threads. Those are the two *worst* situations
to have when one is trying to optimize the
strength of the joint.

Each of those will substantially reduce the
load carrying ability of a connection like that.

A proper fastener will have a) rolled threads,
with b) a bit of a radius or flat at the root.


The power was out here for a few hours, so I did a bit of
reading and playing after cleaning up in the dark.

First, like you, I recalled the NPT thread as sharp, which
is not the case. It's truncated, though obviously not
nearly as much as the UN thread. Neither requires, though
both permit, a radius in the root.

Socket caps over 2", and nonstandard lengths, have cut
threads. Even socket cap screws are allowed to have certain
"discontinuities" (laps and seams) in the thread, though
not below the pitch diameter.

For a ductile material like low carbon steel, stress
raisers have little effect on static strength. They do have
a marked effect on fatigue, though the stress
multiplication factor used for figuring fatigue (Kf) is
often less than for static stress concentration.

The endurance limit for carbon steels varies from 25 to 75
ksi.

Using the 3/4 pipe from before, and assuming a Kf of 3, the
joint should not fatigue below about .138 x 25000 / 3 =
1150 lbs. The 25 ksi is pretty conservative. Kf=2~3 seems
typical for cut threads in low strength fasteners. The
allowable load could be higher if you could predict the
number of cycles and/or avoid strain reversals.

Failure under static load would be closer to .138 x 60000 =
8280 lb.

While the lights were still out, I did a quick experiment.
I threaded a length of 3/8 pipe from the scrap bin thru a
hollow 12 ton porta-power cylinder and put some heavy
washers over the pipe, then screwed a couple random CI
fittings onto the ends. Upon pumping up the cylinder,
initial yield occurred at about 5400 lb tension; tensile
failure was at approx 6800 lb. The failure was exactly
where we both predicted, where the threads exit the
fitting. This corresponds to 57 ksi and 72 ksi
respectively. The material spec for black pipe is 35 ksi
yield and 60 ksi tensile, minimum.

Ned Simmons

Machinest Handbook lookup request (Pleeeeze)
 

In article , Ned Simmons
says...

First, like you, I recalled the NPT thread as sharp, which
is not the case. It's truncated, though obviously not
nearly as much as the UN thread. Neither requires, though
both permit, a radius in the root.

Here of course the *peaks* of the male thread, and
likewise any part of the female thread, don't count.
What only counts is the most highly stressed point,
which is the root of the male thread.

Socket caps over 2", and nonstandard lengths, have cut
threads. Even socket cap screws are allowed to have certain
"discontinuities" (laps and seams) in the thread, though
not below the pitch diameter.

So there is likely to be a fair amount of variation
in the thread quality. That makes sense.

For a ductile material like low carbon steel, stress
raisers have little effect on static strength.

There's a lot of substance in that last statement,
and I'm not sure I understand all of it. My impression
is that in any kind of material, including low carbon
steels, microscopic tears and fractures do reduce the
static strength. Your statement contradicts that
impression and I would be interested in following the
reasoning behind it.

They do have
a marked effect on fatigue, though the stress
multiplication factor used for figuring fatigue (Kf) is
often less than for static stress concentration.

Here I was thinking of the two strength-reducing
mechanisms (stress risers, and the sharp-V stress
concentrating feature) as being two separate but related
mechanisms. Maybe not. But I do know that stress
concentrations from non-optimal geometry - like sharp
inside corners or sharp V thread forms *do* reduce
the ultmate strength in all kinds of materials.

I found this out when making some bolts out of
SP-1 Vespel. :)

The endurance limit for carbon steels varies from 25 to 75
ksi.

Using the 3/4 pipe from before, and assuming a Kf of 3, the
joint should not fatigue below about .138 x 25000 / 3 =
1150 lbs. The 25 ksi is pretty conservative. Kf=2~3 seems
typical for cut threads in low strength fasteners. The
allowable load could be higher if you could predict the
number of cycles and/or avoid strain reversals.

Here I would try to say that the factor of 3 I was proposing
was present even in static load. That's the sort of number
I've seen in practice. Not a fatigue factor, but due
purely to stress concentration.

Failure under static load would be closer to .138 x 60000 =
8280 lb.

While the lights were still out, I did a quick experiment.
I threaded a length of 3/8 pipe from the scrap bin thru a
hollow 12 ton porta-power cylinder and put some heavy
washers over the pipe, then screwed a couple random CI
fittings onto the ends. Upon pumping up the cylinder,
initial yield occurred at about 5400 lb tension; tensile
failure was at approx 6800 lb. The failure was exactly
where we both predicted, where the threads exit the
fitting. This corresponds to 57 ksi and 72 ksi
respectively. The material spec for black pipe is 35 ksi
yield and 60 ksi tensile, minimum.

Did you apply any correction for stress concentration
to go from the 6800 lb to get to the 72 ksi number?
If not then it may be that either a) it's not as big
as I recall, b) the threads you are using are well
formed (rolled?) or c) maybe the black iron pipe was
steel?

Interesting experiment. It would be easy to duplicate it
for some home depot cut thread 3/4 inch pipe, and give
this guy a *real* number. The failure you observed
where the root of the thread crosses the top line of the
female thread is a classic, this is where all the stress
shows up. One reason why designers take pains to avoid
putting *other* stress concentration features in line
with that surface in bolted-up assemblies.

Thanks for taking the time to do this, and to post up
the results. Fascinating.

Jim

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"jim rozen" wrote in message
...
or c) maybe the black iron pipe was steel?

Isn't it? I thought all black and ganvanized pipe was welded steel
tubing.

Tim

--
"That's for the courts to decide." - Homer Simpson
Website @ http://webpages.charter.net/dawill/tmoranwms

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In article ,
says...
In article , Ned Simmons
says...

First, like you, I recalled the NPT thread as sharp, which
is not the case. It's truncated, though obviously not
nearly as much as the UN thread. Neither requires, though
both permit, a radius in the root.

Here of course the *peaks* of the male thread, and
likewise any part of the female thread, don't count.
What only counts is the most highly stressed point,
which is the root of the male thread.

Agreed. Both the crest and root are truncated, though only
very slightly.


Socket caps over 2", and nonstandard lengths, have cut
threads. Even socket cap screws are allowed to have certain
"discontinuities" (laps and seams) in the thread, though
not below the pitch diameter.

So there is likely to be a fair amount of variation
in the thread quality. That makes sense.

For a ductile material like low carbon steel, stress
raisers have little effect on static strength.

There's a lot of substance in that last statement,
and I'm not sure I understand all of it. My impression
is that in any kind of material, including low carbon
steels, microscopic tears and fractures do reduce the
static strength. Your statement contradicts that
impression and I would be interested in following the
reasoning behind it.

And I can't claim to completely understand either.
Presumably it only applies to ductile materials because
they have the ability to deform locally, thus reducing the
stresses without fracturing.


They do have
a marked effect on fatigue, though the stress
multiplication factor used for figuring fatigue (Kf) is
often less than for static stress concentration.

Here I was thinking of the two strength-reducing
mechanisms (stress risers, and the sharp-V stress
concentrating feature) as being two separate but related
mechanisms. Maybe not. But I do know that stress
concentrations from non-optimal geometry - like sharp
inside corners or sharp V thread forms *do* reduce
the ultmate strength in all kinds of materials.

I found this out when making some bolts out of
SP-1 Vespel. :)

I think ceramics would be an extreme example of this. I've
never used Vespel, isn't it similar to Ultem? If so, I'm
not surprised it would be sensitive to sharp threads. I'd
expect something like nylon to be less so.


The endurance limit for carbon steels varies from 25 to 75
ksi.

Using the 3/4 pipe from before, and assuming a Kf of 3, the
joint should not fatigue below about .138 x 25000 / 3 =
1150 lbs. The 25 ksi is pretty conservative. Kf=2~3 seems
typical for cut threads in low strength fasteners. The
allowable load could be higher if you could predict the
number of cycles and/or avoid strain reversals.

Here I would try to say that the factor of 3 I was proposing
was present even in static load. That's the sort of number
I've seen in practice. Not a fatigue factor, but due
purely to stress concentration.

Determining stress concentration factors appears to involve
a lot of voodoo and empirical formulas. Most of the values
I found for the stress concentration factor for fatigue
seemed to come from experimental results and were a
function of material as well as geometry. Roark has pretty
extensive formulas for simple elastic stress concentration
for various geometries.


Failure under static load would be closer to .138 x 60000 =
8280 lb.

While the lights were still out, I did a quick experiment.
I threaded a length of 3/8 pipe from the scrap bin thru a
hollow 12 ton porta-power cylinder and put some heavy
washers over the pipe, then screwed a couple random CI
fittings onto the ends. Upon pumping up the cylinder,
initial yield occurred at about 5400 lb tension; tensile
failure was at approx 6800 lb. The failure was exactly
where we both predicted, where the threads exit the
fitting. This corresponds to 57 ksi and 72 ksi
respectively. The material spec for black pipe is 35 ksi
yield and 60 ksi tensile, minimum.

Did you apply any correction for stress concentration
to go from the 6800 lb to get to the 72 ksi number?
If not then it may be that either a) it's not as big
as I recall, b) the threads you are using are well
formed (rolled?) or c) maybe the black iron pipe was
steel?

72 ksi is simply force/cross sectional area at the root of
the thread (6800 lb/.094 in^2).

I just took a look at the threads with a magnifier and
compared them to an NPT gage. They're pretty rough looking,
obviously die cut, though they do appear to have a slightly
wider flat at the root than the gage.

Black iron is a misnomer, it really is steel.


Interesting experiment. It would be easy to duplicate it
for some home depot cut thread 3/4 inch pipe, and give
this guy a *real* number. The failure you observed
where the root of the thread crosses the top line of the
female thread is a classic, this is where all the stress
shows up. One reason why designers take pains to avoid
putting *other* stress concentration features in line
with that surface in bolted-up assemblies.

Thanks for taking the time to do this, and to post up
the results. Fascinating.

I learned a lot myself.

Ned