Thread torque and engagment lenght calculation. 2

[Carlloss]

Hi all,

Apologise if I refresh old questions however I have been looking for similar threads last few nights and could not find anything that would answer my questions. This is my first job as a design engineer and unfortunately I am on my own in entire company so it is imposable to consult my toughs with anybody. Basically I have application (dome pressure regulator, ref to attachment)and I am struggling to get correct pattern for thread calculation.The pressure regulator have threaded bonnet that is tighten against body thread M40 x 1.5. I have been ask will be possible to redesign this bonnet to dome application with 300bar inlet. Below I put step by step my calculation (also attached). I would be highly appreciated for any comets and suggestions especial that is my first thread.
Inputs:
Bonnet/body material: SS316 UTY=205 MPa; UTS=515 Mpa
Sealing diam. = 27mm
Pressure = 300bar (30Mpa)
Thread size = M40x1.5
Torque = ?
Engagement length = ?

Outputs:
1. I have calculated force acting against bonnet. I used sealing dim. behind the O' ring as a worst case scenario.
D=27mm
A=572,55 mm sq.
F=PA
F=17176,5 N

2.I have multiply this force by safety factor 1.3 to be sure that after pressure been applied I still have compression on O' ring
[F][/1]= 1,3xF= 17176,5 N

3.I try to check if M40x1.5 is sufficient for above inputs.
a) Thread can be stressed to 75% of material yield (I used 75% as I saw something similar in some book but I am unsure what is the best practise, plus what should I used UTY or UTS?)
sigma = 075xUTY = 153, 75 MPa
b) Calculated rqu. tensile stress area.
[A][/t1]= [F][/1]/sigma = 145, 23 mm sq.
c) Calculated tensil stress area for M40x1.5
[A][/t2]= Pi/4 ( D-0,938194xP) sq = 1100,74 mm sq.

4. Required torque to overcome F1.
T=KDP
T=0,15x0,044x22329,45= 147,3 round up to 148Nm

5.Length of engagement.
PDnom =PDmax + PDmin / 2 = 39,219mm

[L][/e]= 4[A][/t1] / Pi PDnom = 35,73mm

My questions are:
1. Does my calculations make any sense? Are they even correct?
2. Can I use shorter length of engagement as I can not change thread type, this must stay M40x1.5? If yes how to evaluate minimum length that wont cause any issue?

I would like to apologise also for any grammar/spelling mistakes as English is not my native language. Thank you in advance for your help.



Karol

 

[desertfox]

Hi Carllos

I think your calculations are okay but there are one or two points I've noticed:-

1/ Stainless steel is well known for galling, particularly when both threads are the same material, so its normal to
lubricate the threads prior to assembly to minimize this. So your torque calculation will be affected if you haven't
already considered this fact. The friction factor F reduces with lubricated components.

2/ In the pdf the sketch shows the components and looking at the male thread portion there appears to be a very small
wall thickness below the main diameter of the thread mainly because of the 27 dia sealing ring, I would be concerned
about that small wall thickness which isn't taken into account because it appears that in your calculations you have
assumed a solid tensile area like that of a bolt or screw.

Ah I've just spotted something else, how or what have you based your thread engagement calculation on? whilst it appears okay I'm curious to its origin.

lastly I would like to congratulate you at presenting your problem very clearly particularly as English is not your native tongue.

desertfox

 

[Tuckabag]

I have a comment that isn't applicable to your questions regarding the threads but may be something you have missed in your design.
Do you have some form of seal between the body of the regulator and the "piston carrier" (for want of a better description)?
See my mark up in red.
You'll need to seal these faces else there will be product leaking through the M40 thread.

 

[desertfox]

Hi Tuckbag

Doesn't the 'o' ring seal with the 27 diameter prevent leakage over the threads or have I missed something?

 

[Tuckabag]

Hi Desertfox,

No, there will be pressure both above and below the piston as shown on the drawing.
The pressure up to 300bar above the piston is used in place of a spring to "load" the piston/poppet assembly and control the pressure on the opposite side of the piston.
The 27mm o-ring should prevent pressure from coming through from the bonnet side, but there's nothing to prevent pressure coming through from the "controlled" pressure side of the piston past the "piston carrier".

 

[desertfox]

Hi Tuckabag

I understand now. I thought there was only pressure on one side.

Regards

Desertfox

 

[Carlloss]

Hi all,

Thank you for all yours valuable comments.

Desertfox,
Thank you for such a quick response. I used 'Machine elements in mechanical design. 4th edition.' - Robert L. Mott, P.E. to crate above calculations (section 18) plus 'Mechanical handbook. 27th edition.' to obtain thread values. We use Krytox for threads lubrication.

Tucabag,
Yes you right. I have missed this O'rig seal between piston holder and body. My foul for not providing correct information.

I have spotted another one as well. The thread should be M44x1.5 not M40x1.5, like I have informed previously. Sorry.

Again thank you for your inputs but I am still confused about engagement length and selection of correct values in my calculations. I have backed to work today and dig out more information about this unit.
It will be pressure regulator with 1:1 ratio (basically if I applied 300bar dome pressure then I am expecting to have 300bar below piston, please see attachment). In this case my effective area and load will be based just on A1 (A1- area for 38mm; A2- area for 27mm). Am I right to say that?
My other concern is, can I even based thread calculation on formulas provided for bolts/screws in situation when the threaded components are not solid thru cross section? Plus is it better to based on yield or tensile strength when calculating required tensile area?
I know that the feelings are not the best why to validate design but sometimes you look on design and you know that something missing but you can describe what. Apology if I repeat myself with some aspects.

Thank you.




Karol

 

[Tuckabag]

Again, I can't help re your thread calcs, but if you're machining the reg body I'd leave more meat in behind the male thread on the body. I'd make the piston holder smaller OD to accomodate this. It just looks too light for these sorts of pressures.
I assume the reg is originally designed for lower outlet pressures than 300bar? If so I'd be suprised if the available thread is strong enough for your required design pressure, not saying it wont do it but I would be suprised.
Also, I'd run a backup ring both sides of the dynamic o-ring on the piston.
Just my 2c worth!

 

[desertfox]

hi

With fasteners its usually best to work with the yield stress.
I'll try and post some more info later.

 

[Carlloss]

Thank you all for your response. So far, on spring loaded application this thread withstand 200 bar below the piston but on slightly different design so this can not be used as a comparison. Anyway thank you for your time and inputs.

Desertfox if find time to post more info. I would be highly appreciated.

Karol

 

[desertfox]

Hi Carllos

Firstly it's best to use yield stress or proof stress for the fasteners your dealing with.

Now whilst you can have any thread you desire,M44 is not a preferred size, there are M40,M42 or M45 which are recognised and dimensions for them can be found in BS3643-2.

As mentioned in my earlier post you cannot use the minimum area of a solid bolt when you have screw threads on a cylindrical tube and furthermore,looking at BS3643 you would have very little material left below the minor diameter of the thread to have any confidence in it not failing in service.

Thread engagement apart from being based on the material strength properties it is usually arranged so that the male bolt will fail first before the nut or tapped hole strips it's thread.
To meet that criteria when both the materials exhibit the same material strength it is usual to ensure that the shear area of the female thread is twice that of the minimum tensile area.
Have a look at this site and the table at the bottom of the page this gives various area's of threads including female and male shear area's

http://www.roymech.co.uk/Useful_Tables/Screws/Thread_Calcs.html

http://www.roymech.co.uk/Useful_Tables/Screws/Thread_tol.html

 

[Carlloss]

Desertfox

Thank you for your deep explanation. I will have a look on this site.

Based on what you said about thread feature machine on cylindrical tube and lack of material behind it, can I assume that it would be better to go with flange type bonnet and use X amount of bolts instead of using threaded bonnet?
With flange type bonnet I could selected correct bolts and be more confident that all calculation are ok.

Sorry for being pain but I would clarify just one more thing. When I am using bolts grade 8.8, 12.9 etc. I always based on proof strength to select correct one, however I had small issue with A2 and A4 grades which dose not have proof strength values (or at lest I can not find it). In this case I was planning to base on 75% on yield values. Can I do that way?
Thank you for your time and help. Apologise if some of my questions are quite obvious.

Karol

 

[desertfox]

Hi

On the website below you can find yield/proof stress for A2,A4

http://www.roymech.co.uk/Useful_Tables/Screws/Strength.htm

Using a flange might be a better solution and regarding using proof stress for bolts thats fine but you need to consider the materials being clamped for example, if you clamp two pieces of aluminium together with a grade 8.8 bolt and use 75% of its capacity, you have to check whether the yield stress of the aluminium in the joint as not been exceeded.

desertfox

 

[Carlloss]

Thank you for answering my questions and help. I will keep in mind all those information during design.



Karol