Proper GD&T on sliding "piston" assembly

[prdave00]

I am working on a design comprised of 2 sliding components with a cylindrical machined spring element disposed between compression platens on the 2 sliding members. I am also considering welding a cylindrical housing (tube) to reduce the likelihood of material impingement between the coils. Please see attached file. Now to my query....

I'm in the process of tolerancing these parts to (a) ensure sliding elements maintain an interference fit, (b) control the wall thickness of the cylinder component, and (c) ensure the spring does not contact any housing/sleeve element that might be installed. To this latter item, the spring element could be laser welded in place to the cylinder element to mitigate this concern and I have already accounted for the spring OD growing in compression.

One of my colleagues suggested both a positional and straightness tolerance be applied to the male - female sliding features as shown. However, I am thinking the positional tolerance alone might suffice. Also, to be candid, I further questioned the need to have both when I couldn't find any examples in ASME Y14.5-2009 that utilized both (not to say that this document is the end all be all in terms of what's not shown).

Any thoughts / guidance on this specific dilemma or in general to my overall tolerancing scheme would be greatly appreciated.

 

[prdave00]

I wanted to correct an error I made. I want to maintain a sliding clearance (not interference) fit to minimize wear / friction between the sliding members.

 

[drawoh]

prdave00,

A positional tolerance locates your accurate features to some external feature, such as a mounting point. Your primary requirement to to control the size and shape of your mating surfaces. You need to control the sliding diameters. GD&T specifications like roundness, straightness and cylindricity are only useful if you have relatively sloppy dimaeters. Accurate diameters control all of this.l

Can you undercut your shaft so that it contacts at only two points? If you can do this, your design will be more reliable.

JHG

 

[prdave00]

Thanks drawoh. I interpreted your post to suggest that my tolerance of size of my shaft OD and cylinder ID might be enough. Adding a positional tolerance my not be neccesary unless I'm want to keep the sliding OD / ID features tied back to say the 'major' diameter on the shaft and cylinder components, respectively, which I do so the spring element never overhangs the afforementioned 'major' diameters. It sounds like I need to revisit the envelope my limits of size give me and then determine if I need to either tighten the tolerance or add a straightness tolerance.

Could you also clarify / expand on your suggestion regarding undercuts on the shaft? I envision 2 cylindrical surfaces seperated by a "groove", but I'm not 100% sure you are not refering to longitudinal running undercuts. Hopefully I stated my confusion (?) in a not so confusing way.

 

[drawoh]

prdave00,

Separated by a groove is what I meant. Most (not all) of the time, you want accurate diametral contacts to be short.

JHG

 

[Belanger]

FYI -- assuming that these are rigid parts, there is no need for the straightness tolerances of 0 at MMC. That concept is already wrapped into the diameter callouts. (In the ASME standard, this is called the "envelope principle.")



John-Paul Belanger
Certified Sr. GD&T Professional
Geometric Learning Systems

 

[fsincox]

prdave00,
While I must agree, I find it ironic, the sugestion of the groove. The real reason is that although rule #1 techncally covers you here, the practicallity of the requirement for any substantial length is suspect in the real world, thus the recommendation. This is a main reason I support the ISO assumption of independancy, and feel it is more in line with engineering logic.
Frank

 

[prdave00]

Frank,
Thanks for the input, but could you clarify which recomendation you are supporting: the additional input from my colleague, JP's input that runs contrary to my colleagues opinion, and/or the suggestion regarding sepearting the linear bearing surface into multiple sections.

 

[fsincox]

prdave00,
Sorry both really, I agree with JP and you that technically the straightness callouts you have are redundant due to rule #1 (ASME's acceptance and requirement of the envelope principle).
I also agree with drawoh's assertion that a (groove) relief is often preferred in cases like yours to minimize the effect on manufacturing of the this same rule (holding perfect straightness over a long distance). We would then now invoke the "continuous feature" on the dimension to achieve rule #1's effect and still have a relieved area in the middle where it is not required. I believe that may be what he is suggesting.
Frank

 

[KENAT]

It's not as simple as 'rule 1 doesn't work in the real world'.

Simplistically applied, rule 1 doesn't take account of the fact that 2 parts may not be perfectly aligned in the real world. This is where you need to allow an adequate clearance, or in some cases make use of projected tolerance zones etc. to allow for this.

Also, long sliding diameters will be prone to contaminants jamming them and the like.

Posting guidelines faq731-376

http://eng-tips.com/market.cfm?

(probably not aimed specifically at you)
What is Engineering anyway: faq1088-1484​

 

[fsincox]

Ken,
That is what is implied and is exactly my point. Another reason for the sperad apart land surfaces.
Frank

 

[pmarc]

prdave00,

What other Guys are saying is definitely true - zero straightness tolerance at MMC will not give you anything as long as you are following Y14.5 standard where rule#1 is default condition. If you use ISO stds. as an indication the independency rule is a default one and then this straightness control makes a sense.

But I would like to go back to your OP for a moment:
I am thinking about the datum scheme you have chosen. My idea (IMHO better reflecting functional relationship of the components) would be to do this as follows:
- PISTON - assign smaller diameter as a datum feature B and specify positional tolerance of bigger diameter wrt it. Why? Because the minor cylinder is in mating relationship with the housing hole, not the major one.
- SPRING - assign inner diameter as a datum feature A and also specify positional tolerance of outer diameter wrt it (maybe in LMC if spring wall thickness is important). Inner diameter is cooperating with the smaller cylinder of the piston, not the outer one.
- CYLINDER - again assign inner diameter as a datum feature C and specify positional tolerance of both outer diameters wrt it. If a wall thickness of cylinder is critical for you you should use position control at LMC on the smaller outer diameter.

If you want to ensure sliding elements maintain at clearance fit you just need to compare MMC of piston smaller cylinder (.172) with MMC of a hole (.173). You can easily see that you will have this clearance directly.

To check if the spring element overhangs the major diameters of piston and cylinder you would have to do slightly more sophisticated tolerance stack-up calculations to see what will happen.

I hope I didn't bring too much confusion to the issue.

 

[drawoh]

fsincox,

What is the difference between a +0/-.0001mm tolerance over a length of 100mm, and a hole to be positioned inside a Ø0.001 circle?

Rule #1 as per ASME Y14.5 makes perfect sense to me. I want a round thing to fit very accurately into a round hole. I can apply ± tolerances to my shaft and hole, and get the fit I want, with a minimum of complex specification. The roundness and straightness tolerance kick in when I apply a sloppy tolerance to the diameter. I keep claiming here that machining drawings are the easiest things in the world to generate, because of the accuracy of the process. Most of my tricky GD&T goes on sheet metal drawings and weldments.

Having prepared your drawing, you now have to evaluate if your fabricator can make the part, preferably at a reasonable price. Hence, the diameter relief. Hence the slot on our second dowel pin hole.

JHG

 

[prdave00]

Thanks to everyone's suggestions.

I received clarification from my colleague who thought we didn't design in any clearance at MMC between the sliding elements and therefore having zero departure from straightness permissible at MMC for both features was critical (granted unrealistic). However we both agreed that clearance should be designed in as I had already done. Now to address possible negative feedback from our supplier I have tightened the tolerance on the shaft OD and increased the tolerance on the bore ID, that still gives me a .0005 diametrical clearance at MMC with perfect form. I don't see any need for a straightness call out even though I find it very tempting for some reason (maybe I'm trying to convince myself it's good from a wear standpoint).

[The bigger concern now is how loud or supplier will scream about the bore depth being more than 7x the drill diameter.]

I've changed my datum scheme as shown in the attachment based on pmarc's suggestion. Any reason to provide secondary and tertiary datums for the positional tolerances since coxiality is of sole interest. For example in the cylinder part (top right corner), my colleague suggested making the primary datum "A" and the secondary datum "B" at MMB. I don't think this is necessary. Any thoughts?

 

[fsincox]

You still have a straightness call out on one?

 

[fsincox]

drawoh,
I understand that rule #1 obviously has its supporters.
Frank

 

[prdave00]

Frank,

I realized I left the straightness tolerance on after I posted. I blame my subconscious.

I also noticed that I assigned MMB to datum "B" in the feature control frame when I think it should be LMB. I'm still familiarizing myself with adding modifiers to the datum in order to convince myself whether I need them. It's my understanding that these modifiers are great to determine requirements of hard gauging, but do you need hard gauging to check for coaxiality of cylindrical features using a CMM (which I fully expect our supplier to use otherwise I may have gone with total run out). I guess this is also why I am challenging my colleagues suggestion which should have read: For example in the cylinder part (top right corner), my colleague suggested making the primary datum "C" and the secondary datum "B" at MMB. Why would I do this?

Once again thanks for all the input.

 

[fsincox]

prdave00,
Good one, You have waddled/stepped into another one there. "Is it going to locate primarily on the face or due to the long length of the diameter does that serve as the primary locator?"
I would use the diameter because of the length, but, "what is the right number to switch you ask?" good question. Basically, are you willing to state a perpendicularity tolerance on the diameter to the face as primary that will guarantee it will happen?, is it realistic to expect?, if no, then probably it locates primarily on the diameter. See ASME Y14.5-1994, pg 66, FIG 4-18 for the discussion on this.
2009 will have a similar example.
Frank

 

[pmarc]

prdave00,
Frank is right. Long cylindrical feature can/should be used here as a primary datum. First because it reflects component's function in the assembly better and second it delivers repeatable relationship of the part to the datum axis in a gage (if gage is used).

I assume your drawings at the moment are not complete, but my additional comment to remember would be that if you specify datums C & E you have to make sure they are referenced in at least one geometric tolerance somewhere on the drawings. If you don't do this then these datum callouts make no sense.

 

[prdave00]

pmarc,

I added the datums C & E in the interest of getting feedback on whether I should either:

Do as my colleague suggested (based on what he was taught) and use these faces as primary datums in the FCF and have the cylindrical feature (B or D, respectively) as the secondary datums, which would then require establishing orientation of the cylinders.

-or-

A compromise (?) and use these faces as secondary datums to remove another degree of freedom.

I'd rather leave those datums off as they relate to the functional features shown. I'm really tempted to just go with total run out on both the cylindrical surfaces and bearing faces but am anticipating, maybe foolishly, that our supplier will prefer positional and perpendicular call outs to better utilize their CMM. If that was the case then I wouldn't have to trouble with determining if boundary conditions (MMB, LMB) are appropriate, removing additional degrees of freedom, etc. At least I think I wouldn't.

When my wife goes to bed I've been pulling out the Y14.5-2009 standard for some leisure reading in hopes that I'll have an epiphany and everything will become clear. Until that happens (or my employer kicks in for a training), I am glad for your generous contributions.