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Post by madpatty on Apr 28, 2021 6:46:06 GMT -5
Hi Guys Ever had one of those days when you need to laugh otherwise you'll cry. My Chinese compressor wheel supplier was ready to ship my couple of wheels, so sent me a pic I thought to myself ..............thats looks nice and shiny , big 128mm inducer 176 mm exducer , reasonable exducer blade shape BUT something didn't seem "right" , had me baffled for a minute or so until I checked out the inducer , BUGGER ME , ITS GOING IN THE WRONG DIRECTION Their Spec sheet gave "forward rotation" , I asked and received clarification that this was correct before making the order . Currently corresponding with the manufacturer to rectify the problem, hopefully it will turn out OK . Cheers John That hurtsđ
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Post by madpatty on Apr 26, 2021 18:41:01 GMT -5
Hi Scott Thanks for the extra pics , very informative , we learn so much more from them :-) Is the quadrant at the "bottom" ( 6 0'clock) region ?? Could you please post a pick of the rear end of the flametube please , the rest of the FT is looking OK . Yep , grind the wheel rather than the NGV for extra clearance to keep as much meat in the NGV as possible. The thrust bearing wear is understandable , normally the thrust is always forward , the rearward thrust you experienced was forced onto the "bump bearing" that isn't really designed for it , its only there to provide the endfloat stop. I'll keep checking the pics out for more clues :-) Cheers John Hi Racket. Reading through the comment I understand everything except the part, âthe rearwards thrust you experiencedâ and the âbump bearing â you are referring to. Which is that component you are referring to? Also how did the rearwards thrust come into play? From the pictures isnât it unclear what contacted what first? is it the turbine wheel that contacted the NGV plate first or the thrust collar contacting the brass thrust bearing first? If the thrust bearing had a contact first then it can be an insufficient oil pressure issue. AND if turbine have had a contact first then the front face of the brass thrust bearing shouldâve had more rubbing marks. Just my 2 cents. Regards Patty
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Post by madpatty on Apr 2, 2021 5:08:40 GMT -5
Hi Ron This would be for short duration use , seconds not minutes/hours.................LOL, I couldn't afford the fuel bill :-) ABB use comp rear wall bleed air cooling on their high pressure marine turbos to allow the use of alloy wheels , for short duration use I could even use water spray , either externally or internally , I've even considered water for the turbine wheel to keep its hub temp down at the higher rpm. This would be an engine for exploring "new frontiers" ;-) Cheers John Hi Racket. Interesting. I am wondering from where do ABB get cooling air for compressor wheels. Regards
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Post by madpatty on Feb 24, 2021 21:43:03 GMT -5
Hi Guys.
Testing Update. Took the engine up to 60000 rpm on my homemade steel ball full complement bearings without any problems.
This thing makes so much noise and rumble that it gets so difficult to keep ramping it up without being afraid of all the power this thing is making.
My white paint from one of the compressor blades flew off at higher rpm so couldnât go higher in rpm without proper rpm reading.
Next step is to improve the rpm measurement technique and attempt the last 10% milestone.
Regards. Patty
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Post by madpatty on Feb 22, 2021 12:19:10 GMT -5
Hi Guys.
Plan was to order some right sized SiN balls so that I can make hybrid ceramic ball bearings of my own and then using those to try and make a full power run.
BUT obviously I couldnât wait for the SiN balls to arrive so went ahead and made some âfull complement steel ballâ bearings to try and run the engine upto 50k rpm.
It was a risky territory for steel balls. BUT test run was successful and engine disassembly after the test run showed the bearings without any stress/heat marks and running smoothly as new.
Video below-
Regards. Patty
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Post by madpatty on Feb 19, 2021 19:34:59 GMT -5
Patty, what is the current squeeze film damper configuration? Inquiring minds want to know. Thanks, Ron Hi Ron. I am using an unsealed ended squeeze film damper much like what is used in Ball Bearing turbochargers. Regards.
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Post by madpatty on Feb 19, 2021 12:26:11 GMT -5
Hi Guys.
So after months of research, modelling, analyses and number of wrecked bearings finally I got this beast running.
Current rpm is 42-43000. Startup to running for 7 minutes straight.
The fluid leaking just before the startup in the video is some diesel that got pooled inside the engine when I mistakenly switched on the diesel pump when I connected itâs power.
Waiting on the SiN balls to get some homemade bearings done to get it up to full power.
Regards.
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Post by madpatty on Jan 20, 2021 8:56:48 GMT -5
Hello everyone thanks for the nice greetings, I've been working on my diffuser design, and the first attempt was not very good, I think the vane angle is not aligned with he stream angle so ill modify that and run it again, it took a long time to configure and run the simulation, but I think it gave me good feedback right now I isolated the simulation restrictions to analyze the diffuser itself, here's a picture and a link to youtube, next step just for curiosity i want to see how the flow looks with no diffuser and then an improved one. also i want to balance the geometry in the way that can be machined on a simple milling machine from a cast part, also im planning to use the complete cartridge core for the turbine so the front plate will work as a mounting point for the diffuser. and will rout the inlet and outlet oil ports thru the difuser . im working on the combustion chamber geometry. any idea on size or shape? i want to keep the best streamline possible bt still be able to build it from steel plates. so far this is how its going. thank you every one and have a great day www.youtube.com/watch?v=PfaWa90xp10Nice simulation. Quite a good amount of whirl (tangential component of air velocity to impeller wheel outlet) component still remains after such long diffuser vanes. Maybe the straightener portion of the vanes is not enough to âstraightenâ the flow out. Regards
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Post by madpatty on Jan 9, 2021 13:04:28 GMT -5
Hi Patty,
You need to determine what the brg system stiffness is and from that you cna find the optimal stiffness.
It is neither the KE or PE (elasticity) particularly that kills the bearings...it is the total energy of the system and how that energy is dissipated to ground or it's surroundings. As I said before, when the energy is stored in the shaft as elasticity, there is nothing that we can do to help dissipate it out of the system. But, if we can soften up the bearings, then we get more movement (KE) and since dampers are just devices that oppose this movement, now we can do something to dissipate this energy as heat out of the system. So you can see the trade...if we have a stiff system, we can control the rotor eccentricity near the bearings, but this comes at the cost of increased bearing loads (due to high amplification factors). If we soften up the system, we can lower amplification factors and thus lower bearing forces, but this may come at the cost of larger centerline eccentricity.
Two additional points....first, imbalance loads go up with the square of the rotor speed, so getting the modes down to lower rpms, means that when we are moving through the modes, we are doing so at regions of lower energy. Secondly, at lower speeds, the required amount of damping is also lower, so it can help if you don't have a lot of length to stuff a damper in there.
Since you have access to Dyrobes, I would encourage you to play around with the bearing stiffness and see where you can "tune" your systems frequency response.
Good luck!
Chris
Hi Chris. How do you determine the bearing system stiffness? One idea is to use the experimental data of at what rpm my bearings are failing. A critical map analysis can be done and rpm vs stiffness can be plotted with critical speed mode lines and then by looking at which stiffness aligns with my experimental critical speed gives me the stiffness. Other question is what control do we have on bearing stiffness. I have assumed 4E6 N/m for my bearings. How can I make them more stiff (at-least 1 order of magnitude more so that critical speeds are above my running speed) when they already are located in a stainless shaft tunnel. Regards.
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Post by madpatty on Jan 6, 2021 23:44:07 GMT -5
Hi Patty Just out of curiosity I've been looking at turbo ball bearing cartridges , the GT55 has a 16 mm shaft and a 44.1mm OD cartridge , similar to the one at the bottom of this page otsturbo.com/products/ceramic-ball-bearing-cartridge/ , now the turbine wheel on the GT55 is kinda small compared to our wheels at only ~111/102 mm , so undoubtedly considerably lighter and less able to generate the loads you are encountering ............there is a very good chance your bearings are a tad marginal , especially at the turb end Cheers John Hi Racket. I am also inclined towards the bearings being marginal. BUT the compressor bearings is what I seems is critical than the turbine bearings. I also intuitively used to think heavier the turbine, more the load on the turbine side bearing which is true up to an extent, but it doesn't explain the entire picture, if you don't take unbalance response and critical speeds into the account. That's what I have been seeing in my testing and my analysis. Here are a few important points that I observed- - Adding a sleeve to increase the effective shaft thickness does decrease the forces acting on turbine bearing BUT it doesn't have much significant effect on the compressor bearing. (Adding a 26mm OD sleeve reduced the turbine bearing force from 9500N to 7900N BUT compressor side bearing forces from 4900N to 4500N)
- Also reducing the unbalance on the compressor side doesn't have any significant effect on the compressor side bearing forces
- Changing the compressor bearing overhang also doesn't have very significant effect on the compressor side bearing forces.
Take into account the unbalance for this is 90g-mm on turbine side and 30g-mm on compressor side so compressor already has 3 times less unbalance than turbine.
I think what's happening is, since in our rotors turbine is much heavier than the compressor, the COM is located very close to the turbine bearing and in conical mode where my bearing failures were concentrated the rotor wants to rotate about that COM(center of mass) thus turbine bearing intrinsically has less motion than the compressor motion (Much like a see-saw with heavier one side and fulcrum also shifted to that heavier side)It's nicely explained in this paper as well-dyrobes.com/wp-content/uploads/2019/02/failure_analysis_of_2_l_engine_turbochargers_gunter-ver2_linked.pdfI am still collecting more data from testing and analysis. BUT in my last multiple tests turbine bearing has survived perfectly and compressor bearings has seen quite some damage. Regards.
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Post by madpatty on Jan 5, 2021 23:14:22 GMT -5
Patty,
Are you still trying to use o-rings as the means to soften the bearing stiffness? Those pictures made it look like the o-rings had taken a set. If so, you might want to go to Parker's site, they have a great engineering resource for o-rings that gives really good guidance on setting the compression and gland depth.
I am working to design a squirrel cage mount for the bearing and SFD in my engine, as the means to soften up the system. I am trying to move the first rigid body mode down and reduce the amount of elastic energy that goes into the shaft, which I can't do anything with. If we can get the energy transferred into the bearing/damper (as KE), then I can design the damper to do something with that. As stated previously, if we can move the mode down to a lower rpm, then we have to deal with lower overall system energy. Of course, we will likely drag the first bending mode down well into the operating range, but with the soft mounts, we will be putting less bending strain into the shaft (and thus less elastic energy into the system). This approach is working well for my me in my model to get the bearing loads reduced to acceptable levels.
Here is my current model...not as nice as Dyrobes, but it is free and it forces you to learn the actual background theory and mathematics. Now of course it goes without saying that it is a linear model and thus there are no cross-coupling effects included.
Here is the Campbell diagram... you can see that the first rigid body mode, 1 per rev crossing, is around 14.5 krpm with the first bending mode occurring around 37.3 krpm...then we are clear. The backward whirl of the second bending mode can be seen slightly at around 74 krpm, but that will not drive the rotor to instability, so we can essentially ignore that one...
Here are the mode shapes at 75 krpm...(obviously, we are not in resonance with any of these at this speed)
Finally, here is the bearing deflection over the speed sweep from 0 to 80 krpm (in Hz, not RPM)
The first rigid body mode crossing drives the bearing forces to around 7 lbf, and the first bending mode is around 23 lbf and then walks back down to around 9 lbf. This is still a work in progress, so I may put up more later when I get the design more firmed out.
Also, I see that Dyrobes can do all the calcs for the SFD, are you simulating them in your models?
Good luck!
Chris
Hi Chris. Thanks for sharing your analyses. I am sort of confused at this point on whether I should use end seals(O-rings) on the damper lands or just use a leaking end damper(like in a BBT). If you'd have any resource where I can get some insights on benefits of one over the other then please let me know. I know the pressure profile will be different for both but how much it effects damping is the question. I am willing to drop the O-rings anytime just to reduce the complexity. Is it the elastic energy that kills the bearings or the KE? Also I see your bearings are experiencing very little force at all modes. Is this because of the soft mounts you are using? I haven't yet tried dyrobes for any damper calculations but was basing my rough hand calculation from a NASA paper by Dr. Gunter. Really straightforward approach to design a damper. ntrs.nasa.gov/api/citations/19750007925/downloads/19750007925.pdf?attachment=trueRegards.
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Post by madpatty on Jan 5, 2021 20:46:27 GMT -5
Hi Patty Have you considered simply fitting a couple of bigger bearings ?? Cheers John Hi Racket. Yes. Bigger bearings with dampers, so that anything and everything is taken care of. Cheers.
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Post by madpatty on Jan 5, 2021 3:51:24 GMT -5
Hi Racket.
Damper length changes by 1/(viscosity)^1/3.
For diesel fuel which is approx. 10 times less viscous than 15W-40 at room temperature, I will need almost 2.15 times damper length to get same damping. Quickly running the Cals for my setup, I will need a damper close to 40mm in length if using diesel, sorta risky.
On a side note, I read that paper multiple times, the viscosity used in that paper for SFD lube is same as 15W-40 at 60 degrees Celsius.
Regards.
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Post by madpatty on Jan 5, 2021 2:38:47 GMT -5
Hi Racket. No problems. I'll figure out something. I am pretty skeptical at using fuel for damping due to it's very low viscosity but it's worth trying. Also what were your bearing positions on the turbine shaft? Same as where brass bush journals are? The 74mm OD of the shaft tunnel for your 9/94 engine is quite tight keeping in mind that you used 47mm OD bearings that too mounted in bronze cups. How did you channelize the oil inlet and outlet? The bearings you used are quite bigger than mine. I think mine are just marginal for the loads posed by this big rotor. Regards.
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Post by madpatty on Jan 4, 2021 2:52:44 GMT -5
Hi Patty It was a long time ago , most pics would have been on the old Yahoo Site , I didn't keep much info on the 9/94 as it was a pretty fast build and the 10/98 was in the pipeline before I'd really finished it . I'll have a look through the CDs I have to find whats here . The O'ring grooves ............if I remember correctly I had problems getting just the right depth so that there was a firm push required on the cup to fit it into its recess, but not too firm that it overode the preload . Do you intend using a "cartridge" or individual cups ?? Cheers John Hi Racket. Cartridge will give more damping so I am leaning towards that. Maybe I can get away with O-rings this way as I will already have a pretty long damper and not have to worry about O-rings on hotter side as well. O-rings like you said is critical in our case. Old yahoo site is gone with all the data as well. Regards.
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