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Пікірлер
@ThePowerhound 6 күн бұрын
Nice video. The only thing I have to say about it is that a size dimension only exists when it’s defining a regular feature of size. It’s not about whether or not it has a +/- tolerance on it. The dimension that controls the step is simply a dimension, not a size dimension.
@brentlackey8316 9 күн бұрын
These are the four fundamental elements. A singular thing is. Multiple things are.
@brentlackey8316 9 күн бұрын
There are self driving cars, but I'm not going to be one.
@brentlackey8316 9 күн бұрын
New York is on Canada's Southern border with the United States.
@brentlackey8316 9 күн бұрын
Where did you get the idea that decreasing the thickness tolerance would improve flatness?
@Gdandtbasics 5 күн бұрын
Decreasing a size tolerance, decreases the amount of form error allowed due to Rule#1 ASME Y14.5 and the envelope priniciple for the ISO GPS standards.
@shawnegan2228 9 күн бұрын
nice video
@Mateusz_16425 10 күн бұрын
@Gdandtbasics :)
@abdeljelilful 10 күн бұрын
But the datum C is not the mid plane of the part! Why you translated it to the mid plane?? Many thanks
@Gdandtbasics 5 күн бұрын
Datum C is the midplane of the part since the datum feature symbol is directly in line with the size dimension. This changes the interpretation of the datum to be a feature of size rather than the surface that the extension line originates from. This rule is outlined in the standards.
@SandeepSingh-43 11 күн бұрын
Can you please clarify how datum C is defined as the mid plane. It seems to defy how Datum’s are represented. Thanks
@Gdandtbasics 5 күн бұрын
Datum C is defined as a midplane since the datum feature symbol is directly in line with the size dimension. This is a rule within the standards.
@greggraham2653 3 күн бұрын
Given that Datum C is the midplane: Datum A remains the primary datum, setting the Z = 0 plane, as it represents the main flat surface on which the part would rest. Datum B is still the secondary datum, providing a perpendicular reference plane, which likely sets X = 0 or Y = 0, depending on the orientation. Datum C, as a midplane datum, will establish symmetry about the part's center. This would mean the Y = 0 (or X = 0, if oriented differently) is located at the centerline of the part based on the dimensioned width. With Datum C as a midplane, the (0,0,0) origin will now be located at: The intersection of Datum A (primary datum surface), Datum B (secondary datum plane or surface), and The centerline/midplane established by Datum C. This zero position would likely be positioned at one of the lower corners of the part in the X and Z directions (as defined by Datum B and Datum A), while the Y = 0 axis would pass through the centerline of the part due to Datum C.
@SandeepSingh-43 2 күн бұрын
@@Gdandtbasicsthank you. I will check the standards.
@SandeepSingh-43 2 күн бұрын
@@Gdandtbasicsso does the True Position of 1 get split?
@marcolima89 19 күн бұрын
As a mechanical engineer in Europe, I really appreciated this video about the ISO standard. Thanks
@radon3292 21 күн бұрын
Why does a profile have B datum for secondary? It makes no difference
@Gdandtbasics 5 күн бұрын
You're absolutely correct, none of the features that utilize this profile callout need B as a datum feature, however its still considered good practice to include it just incase that UOS gets applied to a feature in future revisions that DOES utilize it. Basically it doesnt hurt to include it and make all the feature control frames have the same DRF.
@sjsj4998 22 күн бұрын
What is dim. 35 and 18 limits?
@Gdandtbasics 5 күн бұрын
These are defined by the unless otherwise specified profile of a surface tolerance found in the title block. Essentially creating a located ± 0.375 size dimension for these two dimensions.
@MichaelMachado2 22 күн бұрын
I’ve had the exact same question myself! Thanks for answering.
@Mateusz_16425 23 күн бұрын
Timestamp: 5:32 => The Limits of Size are applied on diameter. Circularity tolerance defines tolerance zone of two concentric circles offset from each other radially by the tol. value. If LoS were applied on radial basis, it would be 1.0010 to 1.0025. By my understanding, it means that LoS define Circularity of .0015 (instead of .003). Please correct me if i this is wrong.
@Gdandtbasics 5 күн бұрын
The limits of size for a diametric feature should always be applied on a diametric basis and not radial. Rule #1 for a cylinder will always result in form control (cylindricity and circularity) that matches the total size tolerance for the diametric feature.
@Mateusz_16425 5 күн бұрын
@@Gdandtbasics Please have a look at Fig. 5-10 in Y14.5-2009, especially the "Means this" section of that figure. It shows that the Circularity tolerance is defined by two concentric circles. These circles ar offset from each other radially by the tolerance value. If we calculate the Circularity from the Limits of Size of the example in the video, it would be (2.005 - 2.002) / 2. This gives the result of Circularity equal to .0030/2 = .0015 (half the value mentioned in video).
@chrmeiss67 24 күн бұрын
It‘s very simple: The two sizes aren‘t based on features function.
@adriandelolmo4693 24 күн бұрын
Hello, first of all, thank you for your work in your videos, they are very classifying. Now we go to my question, I have a part of sheet metal, 1.6mm thick, in which the only important thing is the relative position of 4 holes, the rest of the part is covered with the general notes of the part list. Can I dimension the four holes with a position tolerance without any Datum? I just want to control the pattern between them. In Asme. Thanks
@Gdandtbasics 5 күн бұрын
Yes you can do this, however, most often we see the surface that the holes are on being identified as Datum A and then the pattern of holes is identified as a datum feature with position to datum A. This essentially controls location of hole-hole and perpendicularity to datum A.
@changw_w9413 26 күн бұрын
great video❤
@regaladosuarezguillermo7345 27 күн бұрын
Thank you very much for the video. Could you explain the same three concepts but with the ASME standard? Thank you very much in advance. Greetings from Tijuana.
@Hobz79 29 күн бұрын
I wouldn't read too much into the drawing, it is just an imaginary example and we don't know anything about the other parts it mates with. The designer would have done a tolerance analysis to make sure the parts still fit. The ID could definitely have been datum B, with the bolt pattern or keyway datum C, but without knowing the design it's impossible to say which way is correct
@keydi98 28 күн бұрын
To have less références maybe, B embed an axis and à plane we can use to position anything else.
@patientestant Ай бұрын
Informative. Most of the time lathe operations are perpendicular and the tolerance is wide enough that the ambiguous call out is not an issue. If you are talking about tighter tolerances and on the edge of a tolerance, then this can be relevant, but if you have a relatively wide tolerance then you are right parts can look horrible and still pass, but that often is not an issue, because it is common to know how to avoid having such a large discrepancy. So for simple parts, I am not sure you really need to have datum’s and profile call outs. Once people learn this though, it will help clarify exactly when the parts pass or fail.
@MichaelMachado2 Ай бұрын
Any reason to not just replace any and all size dimensions with profile callouts?
@keydi98 Ай бұрын
That would be ambiguous, all the features are by definition separates it is difficult to join them unless you put an unique tolérance zone by using the UF modificator.... however that would cost too much... for nothing.....
@Gdandtbasics 5 күн бұрын
Thats a great question, and we cover that in our training courses. But the major reason is that profile will control Size, Location, Orientation, and Form all to the same value. Whereas if we control the Location and Orientation with the Position symbol, we can have different size/form tolerances via the size dimension and rule #1. This gives more flexibility to tolerance appropriately to the functional intent of the feature.
@TAH1712 Ай бұрын
There's no point to any 'B' reference. I really do feel sorry for all the machinists that think 'I'm being making great parts that never get rejected - but now I'm not sure I fully understand the drawings anymore !' Buying parts, the specifier has the advantage - shop owners now have to decode the drawing to get back to even the simple features. It is obviously a big competitive advantage to fully appreciate the drawing and identify the dead easy vs the extremely difficult part features. Looking at the drawing it makes no sense to me that the boss isn't datum B with the slot being positioned relative to the bore ID. pattern of holes positioned to the boss and then the slot positioned Looking at the drawing it makes no sense to me that the boss isn't datum B and no sense that the boss ID and OD have differing positional tolerances yet both have the same tolerance of size, how exactly does that 'functionally' work on real parts. The slot position is relative to the bore ID and to make the holes as datum B , the last feature to be produced as a 'functional' datum I still don't understand. Why has the OD flange the same position tolerance with a +/- 0.015 tol as both the boss ID and OD with size tols of +/- 0.001 ? Why has 21 people liked this without anyone questioning anything going on with this part drawing?
@keydi98 Ай бұрын
I had a hard time understanding all your points, the drawing for me is not that bad.... I just wished the ID to be at least a référence..... but Either way when the function requiered are fullfilled it is okay.... By the way the remark concerning the ambiguous linear dimension is true... in Iso gps that type of dimension was prohibed since à lot of time.
@Gdandtbasics 5 күн бұрын
One thing to not forget as a designer, datums should be selected due to their functional importance, which is why the bolt pattern was selected as datum B. It sets the location of this part in the final assembly after datum feature A. Then all other features should be located with respect to this functional location. Thus the key way is clocked with respect to the pattern of holes. The size dimensions should be selected once again from the functional design of the part where a good tolerance analysis will lead you to the appropriate amount of tolerance for all features of size as well as surfaces. And lastly, at the end of the day, this drawing was created to specifically discuss the stepped size dimension rather than all the other aspects of the example.
@TAH1712 5 күн бұрын
@@Gdandtbasics Also, 'At the end of the day' it would really help if simple drawings were complete without any conjecture - as your are fully responsible for your output, I read your last sentence with incredulity - You can not on the one hand, be precise with logic to stipulate unambiguous definition, yet sidestep other illogical tolerances by saying 'this drawing was created to specifically discuss the stepped size dimension rather than all the other aspects of the example.' Sorry - but that's my sincere opinion.
@crashfactory Ай бұрын
GREAT video. thank you!
@jenshavla4673 Ай бұрын
Very informative, thank you.
@edivaldobulba4508 Ай бұрын
The video is very good and everything is correct. We can say that the circularity tolerance zone cannot be fully used as usual because we have to respect the Envelope Principle which means that in the global measurement it cannot exceed 20.5 but in the local measurement between points there cannot be a size smaller than 19.5 while the smallest circle of the tolerance zone is 18.5 Perhaps the main attribute of the Envelope Principle is that it establishes a form tolerance equal to the size tolerance and not only for circularity but for other surface form tolerances such as straightness, flatness, and cylindricity.
@saagusaagu Ай бұрын
Which means we don’t need to verify the bolt circle diameter as long as the basic dim is located. All we are concerned is about the position of individual 8holes. I have one question though, If one of the hole is inspected and measured a dia of .254. What would be the additional tolerance allowed for that hole?
@Gdandtbasics 5 күн бұрын
Your first statement is absolutely correct, we never actually measure the bolt circle diameter. We simply check each instance and its location which indirectly controls the bolt circle diameter. To answer your second question, since the MMC modifier is NOT present we don't get any additional tolerance. However if it WERE present we would get an additional .009 of position to add to the existing .005 resulting in a total diametric position for that single hole of .014 inches.
@xCharjx Ай бұрын
Excellent speech! Well done.
@chrmeiss67 Ай бұрын
Very sophisticated problem. Solution is simple and uses a standard process: - GD&T or ISO GPS of single part from design engineer must use function based datums (= contacts touch or fit against rigid partners). - Derived verification specification must take into account the measurement error and should introduce "verification contacts". - Derived manufactured part specification must introduce manufacturing contacts and manufacturing targets for features that control fit. For castings, datums cannot be used! The contact partner is rigid (casting mould), but the hot metal itself is not rigid! Please see: The function based datums are never replaced, they are not deleted!
@465maltbie Ай бұрын
Thank you. Charles
@luisernestoquiroz1367 Ай бұрын
Great video! Thanks for sharing.
@pardo94 Ай бұрын
God bless you :)
@pardo94 Ай бұрын
Such an interesting video! cheers mate!
@chrisreid8298 Ай бұрын
Im confused by the 10 deg and 20 deg dimensions are basic along with a few others. I do t see any GD&T that contols their angularity so why can they be basic dimensions? Wouldnt that mean that those angles are not toleranced?
@Gdandtbasics Ай бұрын
The 10 and 20 degree dimensions are defining a True Profile. This is the ideal shape of a feature being controlled by a profile tolerance.
@chrisreid8298 Ай бұрын
Big fan of oordinate dimensioning
@Gdandtbasics Ай бұрын
I'm an even bigger fan of Position Tolerancing
@chrisreid8298 Ай бұрын
It seemslike it would be a better practice to design the part to nominal and keep your tolerances symmetric since so many parts are made with CNC, but i'm curious to hear from anyone who disagrees with this.
@Gdandtbasics 5 күн бұрын
In our experience, sometimes people model their CAD at the worse case assembly conditions and then only allow tolerances to remove material. On the flip side, you can model to the "weakest" condition meaning a lack of material for FEA purposes. This would then be a scenario where you want to add material with the tolerance but not subtract any.
@marcolima89 Ай бұрын
great and insightful video as always
@chrisreid8298 Ай бұрын
I have a part that is a spacer plate sandwiched between two flat plates. I dont care how thick it is. +- .0.5 mm would be fine but i want it to be flat. I need to ensure that the plate is flat +-0.1 mm after it is bolted between the plates. The problem is the plastic bows so using a flatness callout of 0.1 is difficult and expensive. How can i communicate what I need on my drawing?
@Gdandtbasics Ай бұрын
Check out the last half of this video here. kzbin.info/www/bejne/nH_WdWyOed9nmLs The flatness he shows on the bottom and the parallelism on the top will restrict the over all form of you flat plate, however still allow you to have a really loose size tolerance.
@chrisreid8298 Ай бұрын
At 7:30 using datum A on the larger flat surface controls 3 DOF and a datum B on the hole controls 2 DOF. Do you still need a third to control the 6th DOF? (I’ll call it the yaw rotation) I just ran into this on a drawing.
@Gdandtbasics Ай бұрын
Another great question. You dont HAVE to. But if there were other features that you wished to control with respect to this 6th degree of freedom with respect to the DRF, you would likely need to rely on Simultaneous Requirements OR identify a tertiary feature, such as the surface being controlled with profile in this example.
@CaynsPayn Ай бұрын
www
@pauldamian2988 Ай бұрын
Good stuff as always. I have a question, though... roughly 10:00 , the two versions shown, WHY are the 2.00 dimension BASIC? Am I missing something silly?
@Gdandtbasics Ай бұрын
Great question! Most drawings have a standard profile control listed in the "Unless Otherwise Specified Tolerances" in the title block. This is sort of a "catch all tolerance" that you can use for some less critical features, however we still need basic dimensions to show the ideal locations of the features. So these 2 inch basic dimensions are identifying the ideal length of the flanges with respect to the datums, and the profile callout will define the tolerance.
@dullvoicereader 2 ай бұрын
Thanks for the video!
@Inna-di-red 2 ай бұрын
Good video
@hakantorstensson8053 2 ай бұрын
Thanks for the clear explanation by moving the lower TZ around! greate
@SoFiAtHeFiRsTaRt 2 ай бұрын
What if the basic dimensions are there for the creation and the profile callout does not have a datum system?
@Gdandtbasics 2 ай бұрын
Thanks for the question! If no datums are listed in the profile feature control frame, then the profile callout is ONLY controlling the form of the surface. However, if the callout is for multiple combined surfaces (consider use of modifiers like between and all around), the relative location/orientation of the surfaces to each other WILL be controlled.
@harikumar4757 2 ай бұрын
Sir, regarding article by GD&T basics on December 21,2014 I have doubt about circularity tolerance,in example (1) regarding dimensional tolerance mentioned the diameter is ∅10.0 with dimensional limit of ±0.08 FOR shaft If it's MMC, ∅ =10.0+0.08=10.08 If it's LMC, ∅. =10.0 -0.08=9.92 The difference between MMC and LMC is 10.08-9.92=0.16 My doubt is ,is this value 0.16 belongs to diametrically or radially. In first example it is shown radially (that is per side) . while calculating Radi of MMC is 10.08/2=5.04( Max Radi) Same way Radi of LMC is 9.92/2=4.96(Min Radi) According to first example limit tolerance is 0.16 So if this limit difference is adding to Radi LMC that is 4.96 +0.16=5.12 That is beyond the Radi MMC(5.04) If adding 4.96+0.08=5.04 My doubt is this 0.16 value ,total limit or perside limit .
@brettkrienke2109 2 ай бұрын
At 8:52 you say when using DMP flatness that those surfaces don’t have any form control. Certainly they are not uncontrolled. Although DMP Flatness breaks Rule #1, the limits of size are still controlling the form on the individual surfaces. There’s just no MMC boundary. Instead there is a virtual condition boundary that the surfaces cannot violate. So there’s still form control. Maybe you just meant it’s not as tight of a control as the “.005” flatness.
@pgabli10 2 ай бұрын
Not true
@ericschertz9110 2 ай бұрын
I'd like to throw in my two cents in the hope that someone reads it and offers a counter opinion. I asked this same question in the GD&T Basics Inspection course and I agree that in an ideal world, anytime the UAME or local size reports out of tolerance, then the part should be rejected. But in my (limited) CMM experience, the UAME and local size rarely report within tolerance on a CMM when dealing with a tolerance that tight unless you have an analog scanning probe or take an extremely dense number of hits. And I'm talking like a point every couple of millimeters (who has time for that?). And even then, it can sometimes report slightly out of tolerance. I'm sure it has something to do with the math involved, which is beyond my knowledge. I've come across this more times than I can count when measuring drill holes for example where the UAME reports too small, yet the plug gauge always fits. But we don't have functional gauges for every single feature that we machine. Sure we have bore gauges and micrometers, but those just give a two point dimension. So I have no other choice but to rely on the average size (or least squares in CMM terms). It's also a lot more repeatable. Do I like it? No, I hate it. But realistically, if I were to fail every part where the UAME or the local size reports out of tolerance on the CMM, then I'd be rejecting about 80% of the parts that we make. And that doesn't help me, the company, or the customer. Again, I want to stress that I get much better results with an analog scanning probe vs taking hit points, but even that has its limits and not everyone has one of those. That also comes with other things that need to be taken into consideration, like using the correct filter for instance. But that's another topic entirely. So in my view, when dealing with tolerances as tight as the one in the video, the average size is often a necessary evil when the UAME or the local size reports out of tolerance. If anyone else experiences the same thing or has a different view entirely, please chime in because I'm still relatively new myself.
@jefffrodermann5348 2 ай бұрын
Eric, I truly appreciate you sharing your experience on this relevant topic. It sounds like you have a great understanding of the manufacturing process and some of the common challenges that are faced when we deal with evaluation methods (least squared, envelope, etc.). I appreciate that you also mentioned being open to more knowledge on the topic. I would say this is the crux of the issue. Having a seasoned metrologist and CMM programmer involved in the measurement process is the key. A common misconception is that it takes more time to scan at high density then to take single points. While this may have been the case years ago, it no longer is. I can capture 8,000 points on my Zeiss prismo CMM in the same time it takes to do a hit pattern of eight single points and 360°. Because of this, scanning always gives us the best chance of accurately reflecting true geometry of the part. The evaluation algorithm, i.e. least squared versus envelope, is determined by the drawing and customer specification. Understanding the end use of the part is key. I agree that it is really the case where both least squared and envelope measure repeatable and with intolerance. However, there are times where envelope is definitely the method to use. This is definitely a topic. I would love to have more conversation about it. Thanks again for your thoughts.
@ericschertz9110 2 ай бұрын
@@jefffrodermann5348 I wish I could always rely on the envelope results from the CMM to pass or fail parts. From a theoretical stand point, it makes sense since it should determine if parts would physically mate together regardless of form error, or conversely if there is not enough contact between parts. But in practice, I would only be passing somewhere between 10-20% of the parts we make if I relied on the UAME and local size results from the CMM because the vast majority of the time, one or both report out of tolerance. I can put a bore gauge or a micrometer on it all day long in a hundred different spots and never get anything close to the UAME or local size, but it will almost always match the least squares (average) results. Even on features for which we have a fixed functional gauge, like a plug gauge for example, the "GO" will always fit, yet the CMM will say otherwise. So I'm thinking that it might be something to do with the math used by the software to evaluate the point cloud or scan data. I think it's called Constrained Least Squares or CL2 Math for short. Or maybe I'm just plain doing something wrong. I'll admit I've only been programming for a little over a year. But I've had over 100 hours of formal classroom training and I scrutinize my measurement routines to death. If I'm doing something wrong, I don't know what it is. At the end of the day, the shop wants to get paid, the customer wants their parts, and I want to have a job. So my hands are tied. I have to go with what gives me the most repeatable results and what agrees the most with the hard gauges. And unfortunately, that's almost always the average least squares result. Will it come back to bite me at some point? Maybe. But I'm almost certain that if I reject every part that fails the CMM's UAME or local size results, I will be shown straight to the door. Thanks for listening. It's nice to vent after a stressful day at work.
@jefffrodermann5348 Ай бұрын
@@ericschertz9110 you have no idea how much I appreciate your comments on this. I've thought about posting an article on LinkedIn about something similar. Envelope is not a one-size-fits-all solution, although it is often communicated as such.
@chrisreid8298 2 ай бұрын
I put GD&T on my drawings to better define my parts and give the machinists more tolerance. But the parts are always quoted more expensive because they have to have more experienced technicians interpret my drawing and they say it requires a CMM inspection now. How would you respond to this?
@Gdandtbasics 2 ай бұрын
Thanks for this comment, unfortunately we hear this very often. And while very complicated parts MAY require use of a CMM to speed up the process, the simple use of GD&T does not inherently implement the need for a CMM. In fact alot of times if the GD&T is truly understood by the inspector, many of the manual inspection processes don't change, only the interpretation of the variable measurements is what changes like in the case of diametric deviations. Long story short, complicated parts require complicated measurement methods and that has nothing to do with GD&T. The first step to alleviating this problem is training on tolerances and GD&T across the industry. We are always happy to help, whether that is through our free content online or more formal training options.
@amirjavadkar1632 2 ай бұрын
Thanks, How to define tolerances for different surfaces on part? For example we have Bumper and we want to define tolerances on this part on different points (X,Y,Z) How we calculate? ❤❤❤❤❤❤❤❤