Needle bearings in a model diesel ?

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Diesel fuel in the engine explodes - it detonates as opposed to simply deflagration burning - this imposes enormous shock loads.
Shock loading doubles stress loads.
Hence considerably larger bearings and stronger piston and crank designs than petrol.
This will be particularly relevant for sizing big and small end bearings.
Plain bearings offer far better dynamic loading via the larger film wedge area being hydrodynamically created (for a given diameter).
It follows that the same hydrodynamic forces created by numerous smaller wedges is going to be less effective.
In either case the lubrication needs to be copious and maintained.
I don't see using needle bearings in a diesel will permit you to get away with less / simpler lubrication.
The needles also need to run on a hardened inner and outer raceways - you do get sleeves for needle bearings but that requires a dismantleable crank design.
Personally, I'd stick to plain bearings.
IIRC, years ago I read a study that found that oil bath wetted plain bearings outperformed both ball and roller bearings.
My 10c worth.
Regards, Ken I
 
This is the bronze big end bearing from my diesel. It’s 25mm bore and 40mm stroke so 20cc. The crankpin is 12mm diameter and the con rod is 12mm thick giving a bearing width of 14mm. There is a low pressure, low volume pumped oil system as it was originally designed as a two stroke. Compression pressure is around 40 bar (600psi). I have no idea of the combustion pressure, it could be 60 bar or more. It runs around 3000 rpm. This has probably had around 1 hour run time plus a lot of cranking with an electric drill during the development.

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Just my two pennoth for what it's worth. I have seen needles used on a big end and have nothing to add in that regard.

My thoughts are around the small end where I cannot see needle rollers working well at all. After all the small end does not rotate very much. Most of the forces would act on the same bottom few rollers all of the time and probably the same roller at point of ignition every time the engine fires. Massive line loading repeatedly in the same place can't be good. I can only see it being alleviated with extra clearance on the shaft to allow the roller cage to gradually roll around, but it would not be a given. A floating gudgeon pin would not help as that could rotate within the piston.
 
I would guess the bearing pressure for a diesel is much higher (as far as non-combustion forces) than a gasoline engine just due to the much higher compression forces.

Just as with a steam engine, the force on the top of the piston is directly related to cylinder pressure (not counting ignition dynamic pressures). If I remember correctly, if you double the steam pressure on a steam engine piston, you double the force on the piston/rod/bearings.

I would think for an IC engine, the same would apply, ie: double the compression ratio and you get twice the force on the components and bearings (check me on that). (Edit: I guess it is a function of pressure in the cylinder, not compression ratio).

I am not sure about combustion pressures, ie: at the point of and after ignition for diesel vs IC engines, so I can't comment on that.
I would guess the ignition pressure of diesel would be similar to the ignition pressure of many combustible liquids such as gasoline.

I am sure there is a lot to combustion dynamics engineering.
A gasoline engine compresses the fuel/air mixture, and then ignites it with a spark.
A diesel gets ignition from the compressed cylinder air temperature, but I am not sure the burn or pressure would be any different than a gasoline engine if you used a spark plug to ignite the diesel/air mixture.
I will have to study combustion dynamics to do more than just make blind guesses.

.
 
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Just my two pennoth for what it's worth. I have seen needles used on a big end and have nothing to add in that regard.

My thoughts are around the small end where I cannot see needle rollers working well at all. After all the small end does not rotate very much. Most of the forces would act on the same bottom few rollers all of the time and probably the same roller at point of ignition every time the engine fires. Massive line loading repeatedly in the same place can't be good. I can only see it being alleviated with extra clearance on the shaft to allow the roller cage to gradually roll around, but it would not be a given. A floating gudgeon pin would not help as that could rotate within the piston.
Having owned several two stroke motorcycles, they all used needle rollers in the little ends and slide fit gudgeons retained with circlips. If this wasn't a good idea I am sure the Japanese wouldn't have used them for decades.
Diesel loading may be different and I have no experience on this.
Rich
 
Thinking back about gasoline, and the dreaded "knock" that can be produced with low octane fuel, I recall additives like lead added to I guess slow down the ignition process, or at least make it a uniform and more controlled burn over a longer span of time.

I don't recall every reading about additives for diesel, or diesel octane ratings though.

From foundry experience, I know that kerosene is significantly more volitile than diesel.

So it seems that folks go to great extents to avoid gasoline engine knock, but for diesels, the knock seems to be accepted as normal.

How do diesels get away with the knock, when gasoline engines don't?

.
 
Green Twin - That "Dreaded Knock" is the IC engine going from ignition to detonation (from burn to explosion) - it wrecks petrol engines PDQ - unfortunately that is a diesel's default operating mode and the engine has to be built to take it.

I have added a file on preignition etc. for petrol / racing 2 strokes etc all about the nasty things that can bump in your engine and why (it's off topic inasmuch as it,s not about diesels).

Regards, Ken I
 

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Thinking back about gasoline, and the dreaded "knock" that can be produced with low octane fuel, I recall additives like lead added to I guess slow down the ignition process, or at least make it a uniform and more controlled burn over a longer span of time.

I don't recall every reading about additives for diesel, or diesel octane ratings though.

From foundry experience, I know that kerosene is significantly more volitile than diesel.

So it seems that folks go to great extents to avoid gasoline engine knock, but for diesels, the knock seems to be accepted as normal.

How do diesels get away with the knock, when gasoline engines don't?

.
For knocking to happen the fuel and air must be premixed. Diesels in theory don't do this, the fuel and air mix only as combustion occurs. In practice there is some premixed combustion because there is a delay between start of injection and the fuel igniting. This fuel does knock (hence the distinctive diesel noise) but there's not much of it so the pressure spike is manageable.

Diesel fuel has a cetane rating. In a sense, it's the opposite of octane rating: a higher score indicates a higher tendency for pre-ignition. Higher cetane rating means less ignition delay, less fuel in the cylinder to knock when combustion starts, thus less rattly noise. You can buy cetane booster additives.
 
Green Twin - That "Dreaded Knock" is the IC engine going from ignition to detonation (from burn to explosion) - it wrecks petrol engines PDQ - unfortunately that is a diesel's default operating mode and the engine has to be built to take it.

I have added a file on preignition etc. for petrol / racing 2 strokes etc all about the nasty things that can bump in your engine and why (it's off topic inasmuch as it,s not about diesels).

Regards, Ken I
Ken-
That is a great article.
I made it through a lot of it, but not all of it.
Very interesting ideas.

I know from a white paper from the head of spray nozzle technology at Delavan that there is a lot to combustion engineering, and things are far more complex than one may suspect.
The Delavan guy said among other things that the hottest flame is not produced by the finest atomized fuel particles, and if the viscosity of the fuel goes up, the flow/discharge from the nozzle also goes up.

Hats off to combustion engineers; that is a very complex subject.

.
 
Re-Thinking the original Needle Bearing question, based on what has been shared so far.

Thanks to EVERYONE for the great contributions. It is amazing the amount of knowledge and inate mechanical intuition on this forum.
So now, I need to re-phrase the question.

The real question at the heart of the thread should have been
"How to make this diesel model engine less messy during operation?"

Yes, I know, diesels by their very nature tend to be smelly and messy, but I seriously want a diesel, and not an air operated engine. Maybe a bit of an oxymoron, if that is the correct term.

The valve train and some other things will need lube, and that probably cannot be avoided. The piston itself, this being a 2 stroke, will need good oil control rings at the bottom. The crank can maybe use roller bearings at both ends.
The crank pin will be a problem.
The piston pin. What about Rulon or Graphite dry-running bearings? If the piston pin is pressed into the rod and the piston has caps to retain the rod in the piston without having to put holes in the sides of the piston, that might work. With a cross-head style piston, (not a cross head style rod and crank arrangement) repair or mods will be easier if there is a complication or failure. Maybe needle bearings, as a mess-free alternative, are not the answer.

More or less thinking out loud. The graphite (Graphalloy) run-dry bearings are intriguing, if they can be made large enough to distribute the load.
Lloyd
 
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Hi Lloyd.
On piston pins (Gudgeon pins in the UK).
They can 2 initial types. Captive or floating:
  1. Captive are a press fit in either the piston or con-rod little end.
    1. Fitted in the con-rod: the plain surface of piston pin acts on the bearing surfaces of the pin holes in the piston. (Favoured by GM small car engines in the 1970s, and various engines since, usually lower performance high volume lower cost engines?).
    2. Fitted in the Piston, often where there is a needle roller or bronze bushed con-rod little end bearing, the pin to piston was a press fit. Actually, as the differential expansion of the piston to pin causes some distortion of the piston when heated, particularly as the piston crown (aluminium) gets considerably hotter then the (steel) gudgeon pin, there is a "bi-metal" type expansion that causes bendin stresses in the piston.
  2. Floating are a sliding fit in both piston and little end of the rod. - Held in place by circlips, or plugs at the ends that simply slide against the bore, without damaging the bore. - Usual for models, due to size constraints?
Crank big-end bearing. There is NOTHING wrong with correctly design needle roller, or plain bearings, but (oil) cooling of the bearing and consequential long life lubrication is necessary for all bearings, including "self-lubricated" designs such as PTFE, Carbon, etc., except short-life jobs. (e.g. Engines fitted to Torpedoes with explosive warheads! They usually had less than 2 minutes running time!). From my design training, the cooling aspect makes the bearings "last forever" (the design life of the product), as a lack of cooling is by far the greatest cause of lubrication failure then bearing failure. Needle rollers have much less heating than plain bearings, that constantly heat oil by "squidging it out" of the gap to generate the hydrodynamic film that prevents metal to metal contact, putting the hydrodynamic film in shear, heating it further. When the oil reaches a "breakdown temperature" in the sheared hydrodynamic film, the film breaks and a sliding metal to metal contact occurs. Adequate oil supply, "oil film strength" (something like surface tension but the "internal tensile strength" of the liquid oil...? - Not being an expert, I can only hint at what this is?) and temperature resistance negate this failure. Needle rollers "in principle" roll across the bearing surface so have no sliding moment at the metal to metal interface.. the roller surface is stationery at the point of contact. But the ends of the needle rollers do slide within the housing and need lubrication, also the needle to bearing surface actually has a microscopic distortion that creates a TINY elastic deformation in the surface so some lubrication and cooling is necessary to prolong long life... A dry needle roller will not last long. (Only 2 miles on my motorcycle after the oil supply was cut-off and the rollers welded themselves to the race and pin!).
Hope this helps your choice a bit more?
K2
 
K2,
Yes, I understand. Lots of considerations (even ones I am unaware of), lots of choices, and some conflicting end-goals. Even though the selection process can be as complex, or as simple, as one wishes to make it, the bottom line is that it all leads to a single choice. Pressure-fed sleeve bearings seem to be the favored choice, but ultimately, each bearing in the engine becomes its own single choice.

The small end of the conrod seems to be the problem child. At one point I had considered a cross-head design to allow a longer stroke to bore ratio, and also eliminating the hot small-end-in-the-piston problem altogether. But that adds another level of complexity to the build, and will make the model noticeably bigger.

Time to ruminate on the choices, and work on properly scaling the Detroit Diesel. The injector and blower are basically complete, so time to design the engine around those 2 assemblies.

https://www.homemodelenginemachinis...ke-will-it-ever-work.31110/page-8#post-393470
 
The small end is subject to some brutal forces, especially in a diesel. I calculated that the peak pressure on my diesel's small end bearing will be approximately 150 MPa, in the worst case scenario of the engine fully burning a stoichiometric charge. That's more than a LG2 bronze bush can withstand without deforming! I'll have to make the small end from phosphor bronze on aluminium bronze.
 
The small end is subject to some brutal forces, especially in a diesel. I calculated that the peak pressure on my diesel's small end bearing will be approximately 150 MPa, in the worst case scenario of the engine fully burning a stoichiometric charge. That's more than a LG2 bronze bush can withstand without deforming! I'll have to make the small end from phosphor bronze on aluminium bronze.
Agreed, 150MPa is very brutal for a bearing in a hot environment.
May I ask what the max cylinder pressure was that you calculated for your engine? Did you also calculate the degrees ATDC when it occurred?
Thx,
Lloyd

edit- P.S. Its like the piston is coasting along for most of the cycle except for when it gets hit with a sledge hammer when the fuel is injected.
 
Agreed, 150MPa is very brutal for a bearing in a hot environment.
May I ask what the max cylinder pressure was that you calculated for your engine? Did you also calculate the degrees ATDC when it occurred?
Thx,
Lloyd

edit- P.S. Its like the piston is coasting along for most of the cycle except for when it gets hit with a sledge hammer when the fuel is injected.
I should preface this by saying that the model I'm using has a lot of assumptions that may or may not be correct, mostly because it has no idea how my combustion system will behave (and to be honest neither do I), so the software just uses default values. Experiments using a test engine and instrumentation would be needed to get good data, I'm only using this as a first pass to help choose part sizing etc. No promises are made that the engine performance will be like the simulation!

With the above taken into account, Lotus engine sim suggests that the peak pressure in this 'runaway fueling' scenario will be 10.2 MPa, 10 degrees after TDC, with combustion continuing until about 90 degrees after TDC. I however took a more pessimistic view: I used the peak pressure expected from a full power run (at a sensible AFR around 25:1, developing about 9.5 MPa in the cylinder) and scaled it up proportional to the stoichiometric fuel flow rate. This way my design max pressure is actually 16.6 MPa, which generates that 150 MPa small end peak pressure. Basically I assume the engine manages to completely burn a stoichiometric charge of fuel in the time it would normally spend burning a normal charge.

The main reason I'm working this way is because the peak pressure depends heavily on when combustion starts (10 degrees earlier causes a 25% increase in peak pressure in my simulations) and how quickly it proceeds. Unless it's told otherwise, the software assumes that combustion starts more or less at TDC, and uses its own approximations for the burn rate. In practice I have no way to know what these parameters will be. So I try to figure out a worst case scenario and engineer the engine to withstand that. It's also not far off taking the loads I actually expect and adding a 50% safety factor, so my two 'pulling numbers from my butt' approaches to the problem are more or less in agreement.
 
I guess your Lotus sim is based on spark ignition and may be slightly misleading. In spark ignition engine the full charge is there ready to be ignited so the speed of combustion and hence pressure rise is based on the flame front speed.

Diesel’s original design injected the fuel over time to control/minimize the pressure rise over the compression pressure. For modern high speed engines this is no more the case but the fuel is still injected over a number of crankshaft degrees, which unless the ignition delay is excessive will limit the maximum cylinder pressure. For full output my pump delivers over about 15° of crankshaft rotation.

The compression pressure will need to be 3 – 4MPa (depending on you combustion chamber design) to ignite normal diesel fuel. The maximum cylinder pressure will probably not exceed 8 – 10 MPa which will ease the small end load. I have a 25mm diameter piston with an 8mm diameter hardened silver steel pin running in a RG7 bronze bush pressed into an unknown aluminium connecting rod.
 
Critical calculations can take many forms when working with a model where a miscalculation might only cause an embarrassment, rather than a career-ending catastrophe. In-depth calculations, modeling, empirical data, scaling, gut-feel intuition. All are good, and with model building, for me, I just want to feel relatively confident with the choices I make. The type of "proof" that would stand the scrutiny of a design review team isn't necessary for me. In retirement, I am finished with that level of stress, ha ha.

My design is based very loosely on the iconic GM Detroit Diesel 2 stroke, particularly the 71 series. It lasted for 60 years in everything from the single cylinder 1-71 to a V16-71. Dirty emissions finally killed it, but the design stood the test of time and engines could go a million miles, with some TLC along the way.

Simply scaling the engine, 4.25" bore x 5.00" stroke, with a piston pin of 1.50" diameter, here is a very simplified approach.
Using a 1/4 scale, piston dia = 1.062" and piston area = .88 sq in vs 14.2 sqin in the original, for a factor of 1/16, so the force in pounds on the piston is 1/16 of the full size engine. but going with a simple 1/4 scale of the piston pin, to .375"dia, taking into account the moment of inertial and the length and deflection, the deflection would be still be 1/4(?) the original, but the force is only 1/16, so scaling the pin should be fine. But the various bearing areas will be 1/16 of the original, too, so the bearing loads in force per area, will be similar to the original. So the bearings will need to be as robust as the originals. But also, going with a pin larger than 3/8" dia would reduce the area loading on the bearings.

Given that the swept cylinder volume will be 1/64 of the original, is it a reasonable assumption to estimate total power as 1/64 of an original 1-71 diesel?

I know these are just kitchen-table napkin calculations, but honestly, if I am not comfortable with these, I assume that something is not right.

Just a lazy approach, I guess, but digging too deep, for me, starts to take some of the fun out of the project. I prefer making chips at the machine, LOL. But we are all different.
Lloyd
 
I guess your Lotus sim is based on spark ignition and may be slightly misleading. In spark ignition engine the full charge is there ready to be ignited so the speed of combustion and hence pressure rise is based on the flame front speed.

Diesel’s original design injected the fuel over time to control/minimize the pressure rise over the compression pressure. For modern high speed engines this is no more the case but the fuel is still injected over a number of crankshaft degrees, which unless the ignition delay is excessive will limit the maximum cylinder pressure. For full output my pump delivers over about 15° of crankshaft rotation.

The compression pressure will need to be 3 – 4MPa (depending on you combustion chamber design) to ignite normal diesel fuel. The maximum cylinder pressure will probably not exceed 8 – 10 MPa which will ease the small end load. I have a 25mm diameter piston with an 8mm diameter hardened silver steel pin running in a RG7 bronze bush pressed into an unknown aluminium connecting rod.
It does have the ability to model a diesel cycle. This is why the peak pressure only goes from 9.5 MPa to 10.2 MPa when I increase the AFR from 25:1 to 14.5:1, you can actually look at the pressure graph it produces and see the pressure takes longer to drop because it considers the combustion duration to be longer at higher fuelling rates.

I just upscaled it in that unrealistic manner because I wanted to be certain that my design had a safety margin for things going wrong, and my brain needs some rationale for the number I choose.

My design also uses an 8mm pin, but my cylinder bore is larger so the loads will tend to be higher. It's good to hear that your RG7 bush is working ok, RG7 seems to be quite similar to LG2 bronze which I assessed to be slightly too weak for my application. I've decided to go with 954 aluminium bronze, of course with a hardened pin. Notably 954 is cheaper and stronger than LG2 or PB1, which helped me make that call! But of course it's way too hard to run against anything less than a fully hardened journal or pin, so the crank and big end bearings remain LG2.

Strangely, a handbook I have on bearing selection also mentions the use of 'duralumin type material' as the small end bearings of heavy duty diesels. I am of course aware of aluminium alloys being used as bearings in small engines with no bushes (just running directly on the alloy) but it's the first place I've seen mention of someone using such a structural alloy as a bushing, let alone in such a heavily loaded application! Makes me wonder if the common and cheap 2011 T3 alloy would be able to perform this function.
 
Makes me wonder if the common and cheap 2011 T3 alloy would be able to perform this function.

1000,
I am sure you have already considered this, but regarding the load on the bearing for the conn rod small end, the total motion of the small end bearing is only about 0.2 revs per one crank rev, with that movement being a rocking motion. This theoretically drops the PV value of that sleeve bearing significantly. That could explain the small end bearings performing better than expected.
Lloyd
 

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