Ben Einstein at Bolt.io also did a detailed teardown of the Juicero, again pointing out how beautiful and totally over-engineered this machine is. He estimates the juice pack is seeing 64psi load in order to get the juice out.

Part of the failure of this whole endeavor is that the machine was so damned expensive! This was an attempt at a "razor and blades" model. Razor and blades is used many places where there's a consumable component; K-cup coffeemakers (allegedly the inspiration for the Juicero), actual shaving razors, computer printers, old polaroid cameras, some video game systems. The whole premise is that you give away the razor, or sell the coffeemaker (or printer) at a heavy loss to get people to then buy the very profitable blades (or printer ink) forever.

But if you make the machine too expensive, this doesn't work. And this thing is PRICEY! So many CNC-machined components, massive gears, thrust bearings, tapered roller bearings, dowel pins, etc. All of these things are wonderful and expensive. These are how you build one juicer, for a lab, to test juice bag  production variability. To see this sort of stuff in a high-volume production item...yeesh. Apple gets away with this in their laptops because it's an aesthetic they're going for, and they're recycling their own super-machinable castable alloy of aluminum to bring it down from "totally bonkers expensive" to merely "high end".

​The marketing of the Juicero fixated on how strong the machine was ("Able to lift two Teslas"). Is this necessary? For some historical background, when HP first went all-in on cheapo printers (to get the printers as cheap as possible to start consumers buying ink), one of its leaders stood on an older HP printer and said "The fact that this printer can hold me up means it's too expensive". The Juicero is

To get to why the Juicero provides several tons of crushing force, we need to step back and take a hard look at what the machine is supposed to do.

• ​To crush a juice bag? No. See AvE's video above of a normal human squeezing juice out of the bag by hand.
• To get the juice out of the bag? Yes!
• To get the juice out of the bag in a "reasonable" amount of time? Very Yes.

​
This is where the designers went wrong. Remember, the goal is to get the juice out of the bag. A 330VDC motor running through a reduction geartrain to a massive Acme screw to put 4 tons of force onto a bag of juice, plus another massive plate to react the 4 tons of force on the other side, all adds up to brute force and ignorance. Very forceful brutality, and beautiful, blissful ignorance.

How else might this have been approached, in a clever way?

1. Don't offer a machine at all, and have the customers squeeze the bags themselves. This has been shown to work in videos all over youtube. Redesign the bags if necessary to make them even easier to squeeze. After all, the bags are where the money is. Often the best design is not to design anything.
2. Make a toothpaste squeezer (which is what AvE suggests in his video). At this scale, you could call it a clothes-wringer, which was a mature technology in 1905. The level of complexity goes way down. Again, the bags could be redesigned to work in a clothes-wringer, like adding barcodes to the bags so the machine can "see" how close it is to the end of the bag.
3. Use a miniature hydraulic pump to squeeze a water bladder at 64psi against the bag. A tiny peristaltic pump (like what's inside the water flosser in your bathroom) can get to 64psi with no sweat. "But what about the 4 tons of resultant force, how are we handling that?". Build the machine to be cylindrical and then you're just talking about a pipe, which can handle 64psi very efficiently.
   1. Or as a variation, if 40psi (municipal line pressure) would get enough squeeze, make a fitting to attach a hose for the water bladder straight to the sink and just turn on the sink.

None of this addresses the zeroth question here of "How many people want to pay $35 a month for juice at home?

So you break open the Mil Standard, and start going through the steps. And by "steps", I mean the standard lays out, step by step, how to establish a sampling plan. We choose a verification level of 3, which is the most stringent, and decide to test batches of chili five thousand cans at a time. From Table 1, level 3 with 5,000 cans gives us Sample Size code M.

From there, we choose an Acceptable Quality Level of 0.25 (0.25% of cans can have ratmeat, because it won't actually kill​ the consumer), which by Table 2 means we will take 315 samples and will accept the whole batch as long as two or fewer samples are contaminated. We will reject the whole batch if three or more samples are contaminated.

If you're doing math along with me, you'll probably say "Hey, how does 2 in 315 get us to 0.25%?" And you'll be right that it doesn't. Not directly anyway. In table 5, the Average Outgoing Quality Limit for our inspection plan comes in at 0.44%, which means that if we test every complete batch of cans whose batch samples failed, and then throw away all the contaminated cans we find, we should have no more than 0.44% contaminated. This still isn't 0.25%...

​The missing link is that the whole sampling plan is predicated on there being a quality control system that manages the inputs. In the case of our "rats in the chili" scenario, the quality control system would be a bunch of traps, or a herd of cats. For the sorts of applications the spec was written for, the quality control system would have been based on Mil-Q-9858. Like many older military standards, it's been replaced by a bunch of different industry standards (ISO 9001 for most applications, AS9100 for aerospace, ISO 13485 for medical devices, etc.)

So this is all good if you're the chili producer, and want to keep your chili relatively rat-free. But what if you're a grocery chain, and you don't make the chili yourself, but rather put your name on somebody else's chili? Accepting 2 in 315 won't get the chili factory to do their best job, and you don't know what the factory is doing to keep rats out.

​What you can do is use a more stringent sampling plan, like what's described in Mil-Std-1916 "DoD Preferred Methods for Acceptance of Product". Mil-Std-1916 has the sort of sampling plan you might want if you were buying canned chili mac, and didn't have control over the chili factory itself. Essentially, start with all the sampling plans in Mil-Std-105, but only use the ones that reject the lots if there are any cans with ratmeat in them.

While assembling a flat pack desk, I found this really neat cam lock studs. Most of the connections in IKEA-type furniture uses cam lock screws and nuts. So for IKEA lock screws, the blue part here would be threaded, and there would be a recess for a Phillips head screwdriver on the stud end. This stud, on the other hand, uses a taper to to expand the blue sleeve as it's pulled.

​I don't know whether the radial compressive load into the particle board will hurt anything, but it's gotta be higher than the induced radial load from threads.

​The problems one might encounter in use are probably that these will not pull two faying surfaces into intimate contact, which a regular IKEA threaded stud will do. Provided the surfaces are already in contact, these should work pretty well.

These are instructions for a flat pack desk, and​ there's much to dislike about them. On the left hand side, we see that the unfortunate builder of this desk is instructed to assemble five separate pieces of the desk, then set them aside. In practice, this means the builder will need space not only to build the desk itself, but also space to store the pieces he/she has assembled early. This is on top of the extra therbligs needed to put down, locate, and then pick back up the pieces built prematurely.

Basically this is a pictorial version of the cautionary tale / cabinet-building episode in ​All I Need to Know about Manufacturing I Learned in Joe's Garage.​

​For the curious readers, these weren't IKEA instructions. IKEA does two things very well; instructions, and meatballs.

 What I did was plot the tangent function from zero to 25 degrees, and then plot Leo St. Clair's approximation (.018 x theta), and the percentage error. From the plot, we can see that .018 times theta is within about 3% of the real tangent for angles up to 25 degrees. So there you have it! Up to 25 degrees, you can safely linearize tangent, and all your mental math becomes much simpler.

As I mentioned earlier, Picatinny Rails are everywhere. Scopes, grips, lasers, lights, sights are all made to mount to them now, as they’re the de facto standard on guns. Of course, we can’t about the “de facto” standard without talking about the actual standard, MIL-STD-1913.

Let’s start with the mathematical description of Datum C
On a drawing, Datums are how interfaces are defined. For the Picatinny Rail, the Datum is what is theoretically the interface between the Rail itself and the sights.

Finally on the right-hand side of the drawing, we have the height of the top point of the rail, with respect to the surface below the bottom of the rail. This is called out as a Minimum dimension, which makes sense here; the rail can be as tall as we’d like, but we need a minimum clearance for the grabbers.

Now that we’ve reviewed how the rail is designed, let’s look at how some common mounts actually interact with the rail.

A mathematically ideal mount would have four rollers/cylinders, arranged in pairs vertically so that their tangential contact points with the rail's beveled surfaces were exactly .108 apart. The pairs of rollers would translate inward toward one another, and mate up with “sufficient” force for all in-spec rails.
There are some problems here though. Since the four beveled surfaces aren't controlled as far as their angles, the contact locations for any given side will shift closer together as the angles decrease, and farther apart as the angles increase. The magnitude of this shift will be proportional to the radius of the rollers. We can minimize this shift by decreasing the rollers' radii, but this will increase Hertzian contact stress (the limit here would be a sharp edge, which has theoretically infinite stress prior to the rail surface being dented/damaged). A narrow radius also allows for possible damage of the mount, which will make all this mathematical perfection irrelevant anyway. (For comparison here, think about how hard it is to put a dent into a trailer hitch ball, compared to trying to nick the edge of a knife).

So while we can imagine a mount that would be mathematically ideal for attaching to a Picatinny rail, we can't build one that will work well in practice. It's much easier to define a datum mathematically than it is to physically interact with one; this is a hard-learned lesson for most GD&T users.

I keep talking about a mathematically ideal mount, but why do mathematically ideal mounts matter when we live in the real world? There's a joke about how physicists calculate milk production, and it ends with "first assume a spherical cow”. When we engineer things, we are making simplifying assumptions, which mean that our answers are only as valid as the simplifying assumptions.

1 MOA shift means what over a 4-inch long rail? Tangent of 1/60 deg (1 MOA) is 0.00029, meaning 1MOA shift equates to .0012 inch shift over 4 inches (.00029 x .4). That's one third the thickness of a sheet of printer paper (.003-.005 thick). This means that it takes very little change to move the point of impact when removing and reinstalling a scope.

To make sure the scope goes on the same way every time, we can can do many different things.

We can make sure that the scope mount grabs the rail with the same force in the same way every time. As the scope mount grabs the rail, it will bend slightly. (The rail itself will crush inward ever so slightly as well, but that effect is very minor in comparison.) If we can make sure that the mount grabs the rail with the same force every time, we can make sure the scope mount bends the same way every time, and the scope will be in the same place relative to the rail. Springs can help with this (like the Bobro mount), as can adjustment (like the LaRue mount), or consistent torque on mounting screws (most screw mounts, also the Aimpoint mount with the torque nob).

We can also make sure that the scope mount attaches to the rail in the same spot every time. The rail will be pretty similar in cross-section at different points, but it won't be exactly the same. This means that our methods above of getting the same force will be hindered somewhat, because our adjustment methods won't work as well. The elastic in a fat guy's sweatpants will grab him about the same within certain waistline limits, but it won't work as well if the fat guy loses a lot of weight. (It should also be pointed out that the rails' straightness tolerances are usually such that zero will be lost just by moving the mount forward on the rail.)

Enter the NATO Accessory Rail (STANAG 4694). This is a rail designed to solve the problems of the Picatinny rail, while still being backward compatible with it. First, let’s look at the dimensioning of the NATO rail itself.

The basic shape of the rail looks similar enough. But there are these pesky metric dimensions where we used to have inch dimensions. So below I’ll inch-icize (anglicize?) the STANAG drawing so we can get an apples-to-apples comparison of the dimensions. What we’ll see is that not all of the dimensions are the same.

Now let’s see the Picatinny Rail one more time…

In comparison to the MIL-STD-1913 cross-section, the biggest change here is that the datum is setup differently. We still have .108 basic height for our two sets of point contacts, and the point contacts are still .748 +/- .002 apart from each other. However, now the angled surfaces are controlled with an angular tolerance (45° +/- 20’); recall that for the Picatinny rail, there wasn’t any control of what angle the surfaces were to each other. This by itself solves most of the potential problems of the Picatinny rail system.
The other major change is that the distance from the bottom pair of contact points (on the bottom angled surface) to the top face now has a tighter tolerance. It used to be +000 / -.020, but now it’s +.000 / - .010, which is half of the tolerance. The bottom angled surfaces now also have their own form tolerance (flatness .003 / 4.000), which was not present on the Picatinny spec.
The top surface no longer has explicit form control (the Picatinny spec has flatness .005), but the distance from the bottom angled surface contact up to the top face has a much tighter tolerance, so I'll argue that the NATO rail is tighter nonetheless.
Most of the clearance dimensions are the same. The width of the outer edges of the rail is identical. The clearance from the top surface of the rail to the bottom of the base has increased by .003, which is backward compatible (since it will allow all Picatinny mounts, and then some), but is also so close as to be meaningless in practice (.003 is the width of a sheet of paper). The actual width and virtual condition of the base is functionally very similar; .010 position tolerance, with datum feature shift. The difference is that the Picatinny rail is positioned at MMC, and thus has bonus tolerance available as the width of the base shrinks (creating more clearance); the NATO rail has no minimum width of the base (which also creates more clearance). Minimal practical difference, as I suspect that most NATO rails will be manufactured with much more clearance than would be allowed by the Picatinny spec.

I've had my mill (a Grizzly G0463}for several years now, and I really hate changing tools. I understand the need for changing tools (it's hard to flycut with a boring head, and it's hard to make slots with a drill) but it's a huge hassle with small mills (Sieg X2 and X3-style mills). Even with a full-sized Bridgeport (or equivalent...) it's enough of a hassle that power drawbars are installed on many/most of them in industrial settings.


First, let's walk through what makes the Grizzly such a pain for toolchanges. THERE'S NO SPINDLE BRAKE! The mill comes with a stupid little spanner wrench, which grabs holes on the spindle face. This means that you have to use one hand to loosen the drawbar, another hand to hold the spanner wrench (which wants to back out as most spanner wrenches do) and a third hand to catch your precious endmill once it drops free.

Enough people find this to be a colossal hassle that littlemachineshop  sells spline wrenches, and two different kinds of spindle locks to grab the spindle spline to prevent rotation while tightening or loosening the drawbar. I built my own spindle spline lock, which works reasonably well, rather than buy one of those linked. (I don't have any relation with LMS, and don't get any sort of sweet kickback money if you click the links.)  I made my own spindle lock, freeing up an entire hand for moving tools into and out of the collet.

Even with something as marvelous as a spindle lock, I still had to change collets frequently, which was still a huge pain. Take the tool out, unscrew the drawbar from the collet, put the collet back in the rack, then reverse for the new collet. (Starrett sells a stepped adaptor for its edgefinders just so machinists don't have to change collets to touch off.)

I started looking into the Tormach Tooling System (TTS), which is all based around a precision spud going into a single 3/4" R8 collet. TTS would get me away from changing collets, but I'd still have to grab a wrench and turn the drawbar to release tools. This takes less time, but it still takes time. Tormach's video below explains how the TTS spud works with the R8 collet, and gives some general background for how R8 collets work.

Tormach's own mills have an optional power drawbar, wherein the drawbar is tightened against Belleville washers. The drawbar is released when a pneumatic cylinder presses the washers flat. This system works well, and many people have done similar things (hoss machine's power drawbar is very well-known among hobbyists running Sieg-type mills).

Using pneumatics is fine and dandy, and how many commercial CNC machines work. To do this I'd need to buy an air cylinder ($$) and also an air compressor ($$). This seems like a terrible plan to me, since I don't have an air compressor, and don't really want one just to change tools. They're loud, big, and this would be the only thing I run with it. Maybe someday, when I own my own house and can put a compressor into a soundproof enclosure.

With pneumatics out of the picture, I started looking into levers, and toggle links. In part this was because nobody seemed to have (or want to use) a big enough air cylinder to act directly on the drawbar-everybody was using a lever system to increase the force. I saw what SDM Fab did with a lever drawbar and thought that that was a good path forward. Their lever system has a detent to hold the collet open. The downside for my application was that it looked like a pain to make on a manual mill, and it was way bigger than anything I think I need. I don't have the machine rigidity of a Tormach and thus can't put as much useful force into the tool, so there's no need to grab onto the tool quite as strongly. Also there are numerous parts which will cost me money, and also the mysterious "hydraulic intensifier".

So what I started thinking about was cams. Instead of using levers, a cam will give greatly amplified force in comparison to a lever, and is self-locking in many cases. This gets me away from detents, and also gets the size down.

The general notion here is that the entire assembly (all the colored bits) will slide onto the top of the mill (not shown), and I'll pull down on a lever attached to the cam (blue). The friction between the cam and the hex head of the drawbar will pull the baseplate (yellow) into contact with the spindle splines. Additional pulling on the lever will rotate the cam and drive the drawbar down, compressing the belleville washers and releasing the collet. This design is the culmination of several years of ruminating and sketching; it's pretty simple, doesn't add extra weight to the head (very important for a manual machine), and doesn't put axial load on the spindle (not that critical in the X3-type mills, but very good for other mills).

Much like most fixtures I've ever designed, this thing is not pretty. I'm using all aluminum (6061 that I have handy for pretty much everything, and relatively hard-wearing 2014 for the cam itself, again what I've got handy) with socket head cap screws and dowel pins holding it all together. Dowel pins and screws are a great way to put together fixtures, because they're very easy to analyze. The tensile/clamping forces I need drive the size and quantity of the SHCS, and the shear forces drive the size and quantity of the dowel pins. This is how IKEA does it - screws to clamp, dowels to take shear, and never the twain shall meet.

There are a few risks I see with this design. 

• Bending/deflecting the drawbar while camming open or closed, possibly causing problems. The frictional force is proportional to the downward force so there's probably a critical relationship between spring stack height and the cam to drawbar friction coefficient. I can calculate that if I start to worry too much.
  
• Wearing down the cam surface over time. This may smooth/burnish the surface at first, but getting aluminum shavings down into the works of the spindle can't be good. If needed I could remachine the cam out of something harder (aluminum bronze, as an example) but that would be more expensive too. So we'll just wing it.
  
• Cracking the top cover of the head. It's a casting, and not super thick. I have a hunch that the greatest load will be while compressing the springs (as opposed to releasing them) so if my yellow baseplate extends to the edge I should be good. Plates are really weak in bending, but if you can get the load out to the edge, the edge is much stronger in compression for the same force. This is why we stack the same sized moving boxes on top of each other when we can.
  

I'll get into the nitty-gritty of this design and what led to it in future posts, especially since I just noticed I don't have the right size drills for the dowel pin reamers I'm going to use.

The basic breakdown between coarse and fine threads is whether you want a screw that is as strong as physically possible (fine pitch threads), or if you want a screw that people will have a hard time installing incorrectly and will be cheaper to build (coarse pitch threads).

Crossthreading happens during assembly, if the screw isn't lined up properly with the hole. It's much easier to do this with fine thread fasteners than with coarse thread. So if you're planning on something assembled by a simpleton, or something that will be assembled and disassembled frequently, go with coarse threads. If you're looking for maximum strength, fine pitch threads are where it's at.

As the number of threads per inch increases, the distance that the threads cut into the screw decreases. If we compare the 1/4-20 screw above versus its fine friend the 1/4-28 screw, we see that the minor diameter for the coarse screw is .1894, while the minor diameter for the fine screw is .2064. That's .070 on the diameter, and equates to about a 10% increase in breaking strength for the fine thread. Extra-fine threads get even stronger.

When you're tapping a hole, you walk a fine line in how much thread you leave in the hole. Let's take a 1/4-20 screw as an example. The Holo-Krome recommended tap drill is .204 (#6 size), which leaves 75% of the possible thread. The thread percentage is saying where the tap drill size was on the spectrum of exactly major diameter sized (0% thread) to exactly minor diameter sized (100% thread).  So since the difference between the major and minor diameter is smaller for finer threads, the margin of error in the drilled size decreases. This means you have to pay more attention when drilling the holes, maybe retire your drills sooner, etc. all of which costs money and time.

The forces of tapping also make fine threads more expensive and time-consuming. Taps aren't indestructible, and they have to transmit a tremendous amount of torque through their flutes and into their cutting edges. The more cutting edges that are in contact, the greater the torque, which means the greater the likelihood of breaking a tap. Breaking a tap isn't the worst thing, because you can get them out in numerous ways, some much more expensive than others. Or you can keep from breaking the tap at all if you go really slowly, or cycle between taper, plug and bottoming taps to make all your downward progress. All of this is more time and more expense though.

There are two main ways of making external threads; cutting, and rolling. Thread cutting is most commonly done on a lathe, and is nice because it's relatively easy to setup and change over to different sizes and pitches. The downside is that you're cutting off material, which means that you're throwing material ($$) away, and that it can take longer per screw.  Below tubalcain (mrpete222) shows how to thread on a manual lathe. Threading on a CNC lathe is much the same, only way faster. (Threading on a Swiss-type CNC lathe is again much of the same, only on crack.)

The other way to make threads is to roll them. In any kind of manufacturing, there's a breakeven point for different types of processes, and threadmaking is no different. If you had a company that was doing nothing but making threads, you'd want to look into rolling threads. The advantages for the producer are that they're way cheaper and faster, and the advantage to the consumer is that the threads are somewhat stronger. Below, Portland Bolt has a neat video explaining how roll threading works.

Finally, what's the "R" stand for in UNRC and UNRF thread? Radius. On the male (external) threads, the root of the thread is radiused instead of sharp. This is very important for fatigue, as repetitive stresses don't like sharp corners and other stress risers.

This probably also does a minor bit of strengthening, since the minor diameter is slightly greater.

I took a scrap piece of MDF board (from a super nice knockdown desk) and stuck in a threaded insert, then used a threaded knob to hold on a roll of solder. The actual soldering iron is held stationary using a bunch of wood screws and some zip ties. We'll see how well the zip ties last, but if they're good enough for auto repair they should be dandy in a silly solder board. The final touch is a scrap piece of plywood and part of a paint stir stick to keep the cords from flopping around getting broken.

I still need to add a helping hand and maybe a spot for shrink tubing.

The head didn't have an entirely flat surface on the bottom, which is why you can see it's sitting on two clamp straps. I was probably picking up bosses originally in the casting, though I can't remember now. Fortunately the flat surfaces available were parallel to the bottom of the head, or I'd have needed a machinists jack or two. The block did have two very convenient through-holes, which I stuck studs through and then tightened down with nuts, visible immediately to the left and right of the spinning endmill in the picture. The through-holes weren't in an ideal spot though, so I used regular step blocks (a LOT of them, as you can see) to hold down the back end of the block. I didn't need to use any stud couplers though, because clamping studs are always longer than you need for some reason.

The slot started out at .651 wide, and we needed to open it up to .800, to a depth of 1.50.   I also measured the distance from each side of the slot to the combustion chamber dish and found that the slot was .025 closer to the right head. I was using an 11/16 endmill (.691), so I did some math and worked out that I needed to cut .042 into the right side, and .067 into the left side.

I zeroed the X-axis leadscrew, came over .042, widened the slot .800 deep, then locked the X-travel, took up the backlash, rezeroed the leadscrew, and came backward .109 to finish the job. The tensioner fit perfectly, and the bike's running today as far as I know.