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- Jan 20, 2022
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- 93
Especially on Shared Sump motor bikesI actually notice engine smoothness “kicks in” per se after the oil is good and settled in, once it’s been run for a few days.
Yes we all obsess over our cars
Interesting observation using Castrol Edge Full Syn 0W-16
- Thread starter Hermann
- Start date
I take that point But its not that important. If you feel its better - its is better to you.
Arguing this sentiment with JoelB a while back, the design of experiment to measure the engine noise would be much more involved than what one might casually think . It would take more than a sound level meter set to DB- A weighting to get usable data, wen the experiment is all sorted, the resulting data must be properly interpreted an reported; Data can be misused to disabuse.
Think of audio equipment measurement . Repeatable data required by FTC for power rating and distortion.
THD has been used for decades. Funny thing, it tells you not a wit about the sound of the amp when it is less than a few percent and, any unit with solid design and in good operation operation and not overdriven easily measures less than 2 percent.
Super low distortion was touted as the primary goal for a great amplifier.
Well it appears that the means to that end (global negative feedback application) made the amps sound MUCH worse and caused rise of other distortion that was not measured.
So great specs to advertise, a repeatable test procedure, but measuring and touting the wrong thing.
In life there is good art and bad art, good music and bad music, good science and bad science.
What's funny about this post is that I agree with your points written here.
Your first sentence essentially says (to me) placebo is OK. And I completely agree. If you think it's quieter on a certain brand, great run that brand. No issues from me. The issue comes in when someone makes a statement "Brand X quieted down my engine" as a statement of fact. This is erroneous since none of us know this factually.
I also agree testing this scientifically would be exceedingly difficult due to so many uncontrolled factors. Which is why I have said over and over NONE of us here on BITOG have the ability to prove one oil brand is quieter than another. We can only think they are quieter.
Here are the relevant figures and the accompanying test. The piston has a "thrust" side and an "antithrust" side, resulting from the angle of the connecting rod. The thrust side is where the connecting rod pushes the piston against the wall and the antithrust side is where it pulls it away from the wall during the power stroke. The piston–wall clearance becomes minimum on the thrust side during the power stroke, whereas it becomes minimum on the antithrust side during the suction stroke.
It is interesting to examine force variation in the piston during its motion. In Fig. 14 the forces due to combustion gas force Fₚ and the inertia force
Fᵢ acting on the piston during its downward motion during the suction stroke are shown. These two forces have to be balanced by the force Fᴿ acting on the connecting rod. Even though vertical component Fᵈ on the connecting rod is balanced by Fₚ and Fᵢ, the horizontal component Fₛ has to be balanced by the thrust reaction from the cylinder wall. Viscous friction force acting at the piston–wall interface is given as Fᶠ.
Forces acting on the piston during successive strokes
Viscous force and the clearance (film thickness) variation at thrust side
Viscous force and clearance (film thickness) variation in antithrust side
It is interesting to examine force variation in the piston during its motion. In Fig. 14 the forces due to combustion gas force Fₚ and the inertia force
Fᵢ acting on the piston during its downward motion during the suction stroke are shown. These two forces have to be balanced by the force Fᴿ acting on the connecting rod. Even though vertical component Fᵈ on the connecting rod is balanced by Fₚ and Fᵢ, the horizontal component Fₛ has to be balanced by the thrust reaction from the cylinder wall. Viscous friction force acting at the piston–wall interface is given as Fᶠ.
Forces acting on the piston during successive strokes
Viscous force and the clearance (film thickness) variation at thrust side
Viscous force and clearance (film thickness) variation in antithrust side
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Here are the relevant figures and the accompanying test. The piston has a "thrust" side and an "antithrust" side, resulting from the angle of the connecting rod. The thrust side is where the connecting rod pushes the piston against the wall and the antithrust side is where it pulls it away from the wall during the power stroke. The piston–wall clearance becomes minimum on the thrust side during the power stroke, whereas it becomes minimum on the antithrust side during the suction stroke.
It is interesting to examine force variation in the piston during its motion. In Fig. 14 the forces due to combustion gas force Fₚ and the inertia force
Fᵢ acting on the piston during its downward motion during the suction stroke are shown. These two forces have to be balanced by the force Fᴿ acting on the connecting rod. Even though vertical component Fᵈ on the connecting rod is balanced by Fₚ and Fᵢ, the horizontal component Fₛ has to be balanced by the thrust reaction from the cylinder wall. Viscous friction force acting at the piston–wall interface is given as Fᶠ.
Forces acting on the piston during successive strokes
Viscous force and the clearance (film thickness) variation at thrust side
Viscous force and clearance (film thickness) variation in antithrust side
Look very closely at the study you are referencing. It does NOT address the dynamics or lubrication of the top compression ring in any manner. The graphs you posted above from the study are the film thickness between the piston skirt and cylinder wall. Do you think the film thickness is going to be almost 100 microns on the top ring when all the other studies show way less than that. Look closely at that study again.
Do a search for the word "ring" in the PDF. This is the only thing they say about the ring, except for Fig 6 which doesn't even make sense when you look at the text that references Fig 6.
This is from page 343 of your referenced study:
"Future extension of this model should take into account boundary interactions at TDC and BDC, particularly with much thinner films encountered between the piston ring and the cylinder liner."
I found a very good study that addresses exactly what we are debating, and it addresses the combustion gas pressure behind the top ring and how that effects the lubrication and film thickness on the compression stroke when the gas pressure is very high. Give me a little time to put the info together to somewhat summarize the study in the next post. I will also give a link to the downloadable PDF.
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Ref: https://www.researchgate.net/public...ngines/link/5d9b09d0a6fdccfd0e7f3b22/download
F
Figure 1
On Page 7:
"The combustion gases flow into the crevices between piston ring, piston grooves and the cylinder liner as the cylinder pressure rises. The generated contact pressure for the piston ring is determined by the trapped mass behind the ring (i.e. the control volume B)."
Figures 3a and 3b
From Page 10:
"The piston compression ring undergoes hydrodynamic regime of lubrication for a large part of the engine cycle. Asperity interactions occur at piston reversals (particularly at the top dead centre) due to a reduction in lubricant entrainment into the conjunction. The associated friction indicates that boundary interactions take place at the top dead centre in the combustion (power) stroke (Fig. 3b). Contact pressure is distributed more uniformly for the flexible ring and there is a rise in the minimum lubricant film thickness. Therefore, boundary friction is marginally mitigated with a flexible ring."
Figure 6a
From Page 10:
"Mixed regime of lubrication occurs during piston reversals and especially during the combustion stroke due to the reduced sliding velocity and increased contact load, resulting in the reduction of the minimum film thickness (Fig. 6a)."
"A sudden reduction in the film thickness appears during the maximum cylinder pressure."
======================================================================
My comments: This study seems to be very through and detailed IMO. It specifically addresses the combustion pressure effect on the top ring which increases the ring face force to cylinder, and how that effects the film thickness at TDC power stroke.
Also, the studies I've seen don't seem to address the effect of the combustion on the level of lubrication film on the cylinder wall. The combustion has to burn away some oil film which makes the environment for the top piston ring even more harsh in terms of getting adequately lubricated. I think of of the studies referenced before in this thread make a mention that their study/model didn't account for that "lack of lubrication" factor.
The bottom line is that the combustion pressure has a large effect on the film thickness and lubrication of the top compression ring on the power stroke. This is why the top ring and the top area of the cylinder wears the most (ie, cylinder taper and ring groove at TDC) compared to the rest of the cylinder and the 2nd compression ring and oil control rings. The top compression ring IMO is probably the most abused component in an ICE. Many times, the ring(s) will wear out much faster (resulting a a loss of compression) than the rod, crank or cam journal bearings.
Figure 1
On Page 7:
"The combustion gases flow into the crevices between piston ring, piston grooves and the cylinder liner as the cylinder pressure rises. The generated contact pressure for the piston ring is determined by the trapped mass behind the ring (i.e. the control volume B)."
Figures 3a and 3b
From Page 10:
"The piston compression ring undergoes hydrodynamic regime of lubrication for a large part of the engine cycle. Asperity interactions occur at piston reversals (particularly at the top dead centre) due to a reduction in lubricant entrainment into the conjunction. The associated friction indicates that boundary interactions take place at the top dead centre in the combustion (power) stroke (Fig. 3b). Contact pressure is distributed more uniformly for the flexible ring and there is a rise in the minimum lubricant film thickness. Therefore, boundary friction is marginally mitigated with a flexible ring."
Figure 6a
From Page 10:
"Mixed regime of lubrication occurs during piston reversals and especially during the combustion stroke due to the reduced sliding velocity and increased contact load, resulting in the reduction of the minimum film thickness (Fig. 6a)."
"A sudden reduction in the film thickness appears during the maximum cylinder pressure."
======================================================================
My comments: This study seems to be very through and detailed IMO. It specifically addresses the combustion pressure effect on the top ring which increases the ring face force to cylinder, and how that effects the film thickness at TDC power stroke.
Also, the studies I've seen don't seem to address the effect of the combustion on the level of lubrication film on the cylinder wall. The combustion has to burn away some oil film which makes the environment for the top piston ring even more harsh in terms of getting adequately lubricated. I think of of the studies referenced before in this thread make a mention that their study/model didn't account for that "lack of lubrication" factor.
The bottom line is that the combustion pressure has a large effect on the film thickness and lubrication of the top compression ring on the power stroke. This is why the top ring and the top area of the cylinder wears the most (ie, cylinder taper and ring groove at TDC) compared to the rest of the cylinder and the 2nd compression ring and oil control rings. The top compression ring IMO is probably the most abused component in an ICE. Many times, the ring(s) will wear out much faster (resulting a a loss of compression) than the rod, crank or cam journal bearings.
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There is an effect, yes, but it's pretty small as seen in the figures. However, the effect on the friction seems large.
The ratio of the force applied by the gas pressure onto the ring to the thrust force applied on the ring by the connecting rod will be on the order of:
face area of the ring / top-surface area of the piston / sine of the angle from vertical of the connecting rod
This ratio should be pretty small.
It is a very difficult calculation though. The authors say that it was never done before. I don't know how much I trust their calculation, but it is interesting.
Also note in the figures you posted that the blowby pressure expanding the rings is increasing the MOFT, not decreasing it, as I previously stated. This is probably why the friction is less with blowby pressure expanding the rings, as a slightly higher MOFT can dramatically reduce the friction in the boundary lubrication regime if you look at the Stribeck curve.
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What figures are you referring to?
The friction increasing on the power stoke can also be caused by the ring face being forced more against the cylinder wall from the combustion pressure behind the ring like shown in the Figure 1 I posted above. I'm saying the combustion gas pressure effect on the ring is larger than any forces the ring goes through during a power stroke.
As mentioned more below, the rings float in the piston groove, so how can the side load on the piston cause the rings to see more load when they will move in their ring groove if the piston is forced against the cylinder wall. There might be some very short impulse inertial load, but fact is the combustion gas pressure forces the ring against the cylinder during the whole power stroke, and that happens on the entire power stoke to some degree as shown by the graph below.
Look closely at the slope in Graph 3a (below) of the oil film thickness on the combustion stroke compared to the other strokes. The oil film thickness curve vs piston position on the combustion stoke is different throughout the whole stoke because of the combustion gas pressure effect on the backside of the ring.
Graph 3a from the study.
Look at this study (link below - downloadable PDF) which addresses ring sealing behavior, but it's interesting in this discussion because they show what kind of pressure is produced behind the rings during the piston stokes. Obviously the top ring gets hit with a lot of combustion pressure which forces the face harder against the cylinder wall, which causes more ring friction (and causes reduced MOFT too as shown in the other study) ... that's why the top ring and top of the cylinder wear. Friction means increased wear, which is caused by a combination of higher forces and reduced lubrication (less film thickness).
https://www.mdpi.com/2075-4442/9/3/25
I don't agree, because the fact is that the ring floats in the ring groove. While the piston gets slammed against the cylinder to take up the small piston to wall clearance during a power stoke (actually happens on all strokes), the rings will simply readjust themselves in the ring groove. Once the piston makes that initial impulse tilt and hits the cylinder wall, the piston tilting movement stops. Probably happens in a few milliseconds.
The piston slightly tilting and slamming against the cylinder wall actually happens at TDC and BDC on all stokes. The film thickness reduction you see on the TDC and BDC stokes, besides the TDC power stoke, is the piston dynamics factor on the ring oil film thickness - including the slight piston tilt when it changes direction. Throw in the combustion pressure at TDC power stoke on top of that, and you add more ring face pressure during the entire power stroke as seen in the graph 3a above, especially close to TDC when combustion pressure is highest.
If you could run the engine with an electric motor, you wouldn't see any difference in top ring oil film thickness on the combustion stroke ... all four stokes would look the same on Graph 3a without the combustion gas pressure factor. Through in high pressure gas from combustion and you get what you see in Graph 3a above.
If you can find a study that clearly shows that the tilting of the piston during the TDC power stoke is the main factor why the oil film thickness is lower than all other TDC or BDC events, please post it up.
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First of all, this is a great discussion.
I have not heard of the piston hitting the cylinder wall before. One might think that would cause damage to the cylinder walls within a short period of time. I had always thought the rings were the only contact with the cylinder wall.
Having said that, are the pistons actually hitting the cylinder wall itself or is there a oil film or tribofilm that is there to avoid metal to metal?
There is a film of oil of course between the piston and cylinder wall, it was discussed in the one paper Gokhan posted a link to. Of course the piston skirts hit and rub on the cylinder walls ... look at the wear marks on any piston (example below) and you can see where they rub.
Many pistons have a anti-friction coating on the skirts to help reduce friction and wear, and to help keep the piston clearance more under control. Ever hear of "piston slap" ... that's the pistons actually rocking in the cylinder bore when they are cold and before the skirts expand (some GM engines have this characteristic). The skirts are literally "slapping" on the cylinder wall. BTW - pistons are not really round ... they are egg shaped across the shirt so only the thrust areas of the skirt rub the cylinder wall.
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3(a) and (b).
Look at this figure below. It is a simple static-equilibrium problem. Something must counter-support the lateral force of the connective rod, and this is not an instantaneous force—the angle of the connective rod is nonzero for most of the cycle. If the ring is not counter-supporting, it must be the piston itself, but that would cause more wear I would think. Moreover, this lateral con-rod force is much larger than the force applied by blowby on the ring. My guess is that the ring is almost completely pushed in on the thrust side but floats on the antithrust side.
On the contrary, according to Figure 3(a), MOFT is larger when there is blowby vs. when there is no blowby. This larger MOFT with blowby is resulting in less friction with blowby vs. without blowby if you look at the inset in Figure 3(b).
Re: Comments on your bolded text: 1st one - Of course it's the piston itself hitting the cylinder wall (due to rocking in the bore) that is supporting the piston forces from stroke motion. As said multiple times, the rings float in their grooves. The rings on the thrust side can compress all the way into their grooves to the point where the piston in the ring groove area is rubbing on the cylinder wall, along with the rings. The rings float and don't really support the piston in any way. Piston dynamics is all about the piston rubbing on the bore, as discussed in the paper your figure above came from. 2ne one - Yes, the rings obviously can be pushed into their grooves all the way if the piston tilt is extreme. Watch the video I posted in post #90 - they show piston contact to the wall in the ring area and even above the rings on the circumference of the dome area.Look at this figure below. It is a simple static equilibrium problem. Something must counter-support the lateral force of the connective rod, and this is not an instantaneous force—the angle of the connective rod is nonzero for most of the cycle. If the ring is not counter-supporting, it must be the piston itself, but that would cause more wear I would think. Moreover, this lateral con-rod force is much larger than the force applied by blowby on the ring. My guess is that the ring is almost completely pushed in on the thrust side but floats on the antithrust side.
The gas blow-by is probably giving a gaseous film between the ring face and cylinder wall, which helps reduce friction. Fact is, all this info from the studies referenced show that the oil film is reduced throughout the power stroke when there is no gas blow-by, and it's pretty easy to see that it's from the combustion pressure effect on the back side of top ring as shown in Figure 1 in post #86 and Figure 3a in post #88.
I've seen nothing that says the thinning effect on the oil film thickness during the power stroke is due to the forces of motion dynamics of the piston assembly. Like said in post #88, if you rotated the engine at some constant RPM with an electric motor on the crankshaft, the oil film thickness vs piston position curves would all look the same for all four stokes. Throw in combustion on the power stroke and you then see the effect on the oil film thickness on the top ring due to the pressure pushing the ring harder against the cylinder wall, which reduces the oil film thickness and increases friction and wear.
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So, you're basically saying that the rings carry almost no load and the piston skirt does all the hard work of supporting the lateral force of the connective rod. I would have to see a study to believe that.
Look at Figure 3(a) again. It is contradicting what you're saying. When the blowby is included, MOFT is increased, not reduced like you claim.
Of course the piston rubbing on the cylinder wall it carrying all the side load induced by the piston stroke. Have you ever held a piston assembly in your hands and looked closely at it? Tell me how rings that float in/out of their grooves can support the piston when the piston is up tight against the wall and the rings are pushed all the way into their groove to be even with the piston surface. The rings can't support any more force than beyond their spring tension against the wall ... that's why the piston can tilt in the bore and rub on the wall if the piston skirt to cylinder clearance is excessive, just like discussed in the video I linked in post #90.
I think it's actually Figure 3b, with the little exploded window inside the graph showing more friction detail around TDC. Blow-by gas at the highest pressure at TDC could certainly cause the ring friction to do down because a layer of gas could lift the ring face off the wall somewhat. Figure 3b is showing ring friction, not MOFT. In figure 3a, the oil film thickness is basically the same throughout the power stroke regardless of which ring it is.
Think about it. If there is a high pressure gas blowing by the ring then there must also be a relatively large gap between the ring face and cylinder wall that leaks some combustion pressure - that can basically reduce the force of the ring face on the wall. It was on a "flexible" ring too, so that may allow more blow-by. If oil can occupy that space or not (they don't really say or explain the ring blow-by data) then maybe that's why the film thickness is higher - but they never show a blow-up graph of the oil film thickness, it's a graph of ring friction which can be a function of multiple factors.
Anyway, the bottom line is there is plenty of other information that supports that combustion gas pressure pushing on the back side of the top ring during the power stroke effects the top ring film thickness, friction and wear. I really haven't seen anything that says otherwise.
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I'm loving this thread.
Nothing additional to contribute to the MOFT discussion, but I'll add to the discussion on base oil viscosity. Gokhan is correct in that a 0W-16 can have a higher viscosity base oil than a 5w30. You can make a 0W-16 with straight 6 cSt base oil and no VII. You can even achieve this with >50% group III. A 5w30 will typically have 20-25% of 4 cSt base oil blending with 6 cSt and 3-8% VII. Taking this into consideration, it's plausible for a 0W-16 to have a better HTFS than a 5w30.
Nothing additional to contribute to the MOFT discussion, but I'll add to the discussion on base oil viscosity. Gokhan is correct in that a 0W-16 can have a higher viscosity base oil than a 5w30. You can make a 0W-16 with straight 6 cSt base oil and no VII. You can even achieve this with >50% group III. A 5w30 will typically have 20-25% of 4 cSt base oil blending with 6 cSt and 3-8% VII. Taking this into consideration, it's plausible for a 0W-16 to have a better HTFS than a 5w30.
I think we are both right here. This paper has some data:
https://cyberleninka.org/article/n/583816
For a 75 mm ⌀ × 1 mm ring, an 11 MPa of pressure corresponds to a 2,600 N of force. This force is being applied on the ring from the inside of it by the blowby pressure as you said, and it is the force that is counter-supporting the lateral force of the connective rod as I said (action–reaction principle). Without the blowby pressure, the ring would be entirely pushed into the groove to support the 2,600 N (580 lbf) of lateral force applied by the connective rod.
Now, as I said earlier, the blowby is helping the lubrication by making the MOFT larger (not smaller as you said) by more uniformly expanding the ring than the very weak ring spring would do. Even a small change in the MOFT can amount to a large improvement over lubrication (large reduction in friction) in the boundary lubrication regime.
The MOFT is smaller at firing TDC because the force on the ring is larger at firing. While it is true that this force is equal and opposite to the force by the blowby pressure on the ring (action–reaction principle), even when there is no blowby (red curve in Figure 3(a) in your paper), the same phenomenon happens; therefore, the blowby is not the reason for the large force. There would be the same large force even if the ring were entirely rigid with no blowby because of simple static-equilibrium arguments—the ring does counter-support the lateral force of the connective rod. Likewise, the Shell paper that did not include the blowby reported results similar to those in your paper that included the blowby.
I wouldn't give any credibility to your postulate about the blowby directly lubricating the ring. The oil and the tribological films do the lubrication—not the blowby.
This discussion veered quite off from the effect of the base-oil viscosity and VII on the lubrication of the valvetrain and piston rings, but we've learned quite a bit. I say you should believe that the shear rates in the valvetrain and at the rings can be as quite high as the Shell paper I posted earlier reported. Don't neglect the importance of the base-oil viscosity and the irrelevance of the VII-boosted viscosity, and next time fill in that Corvette of yours with an oil with a high base-oil viscosity such as Mobil 1 V-Twin 20W-50 if you care about your valvetrain and piston rings.
Quick scan of that paper, and a quick comment on the ring blow-by after scanning that paper. They describe the "blow-by/blow-back" as gas that gets past the rings through the back side of the ring pack - see Figure 5. They don't show any reference to the blow-by going past the ring face and cylinder wall that I can see.
So in the study I reference, they don't really describe exactly what the "blow-by" refers to, or show any figure describing exactly what they mean. You would think that if high pressure gas was actually blowing past the ring face and the cylinder wall that the pressure on the back side of ring would be counter acted and force on the back of the ring negated somewhat, which would reduce the force pushing the ring outward, and thereby increasing the film thickness (and reducing friction) because of less ring force squeezing out oil between its face and wall. Any gas blow-by pressure on the face side of the ring will cancel out some pressure on the backside of the ring. If the ring was perfectly sealed between the ring face and wall (zero ring face blow-by gas pressure), then all the combustion gas pressure would act to force the ring face against the wall, thereby reducing MOFT and increasing friction.
Also look at Figure 7 in your link reference (see page 11).
"Calculated combustion pressures acting on ring face profiles are presented in Figure 7. The pressure values are highly influences by ring clearances and also liner deformations. Pressure acting on first ring face is very similar to combustion pressure in a combustion chamber."
Just as I noted above, if the pressure on the ring face is very high right around TDC during combustion, the gas force on the back side of the ring is essentially cancelled for a short time, and that would mean less ring expansion force and higher MOFT between the ring face and cylinder wall. As the piston moves downward, then the pressure above the piston drops off pretty quickly and the ring starts sealing better between the ring face and wall.
Anyway ... good find on that last paper ... will digest in more detail later, and maybe comment more on your post #96 if I find something. In the middle of doing my taxes, and looks like I owe $$.
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It looks like the blowby is flowing behind the rings in their model according to Figure 5.Quick scan of that paper, and a quick comment on the ring blow-by after scanning that paper. They describe the "blow-by/blow-back" as gas that gets past the rings through the back side of the ring pack - see Figure 5. They don't show any reference to the blow-by going past the ring face and cylinder wall that I can see.
So in the study I reference, they don't really describe exactly what the "blow-by" refers to, or show any figure describing exactly what they mean. You would think that if high pressure gas was actually blowing past the ring face and the cylinder wall that the pressure on the back side of ring would be counter acted and force on the back of the ring negated somewhat, which would reduce the force pushing the ring outward, and thereby increasing the film thickness (and reducing friction) because of less ring force squeezing out oil between its face and wall. Any gas blow-by pressure on the face side of the ring will cancel out some pressure on the backside of the ring. If the ring was perfectly sealed between the ring face and wall (zero ring face blow-by gas pressure), then all the combustion gas pressure would act to force the ring face against the wall, thereby reducing MOFT and increasing friction.
Also look at Figure 7 in your link reference (see page 11).
"Calculated combustion pressures acting on ring face profiles are presented in Figure 7. The pressure values are highly influences by ring clearances and also liner deformations. Pressure acting on first ring face is very similar to combustion pressure in a combustion chamber."
Just as I noted above, if the pressure on the ring face is very high right around TDC during combustion, the gas force on the back side of the ring is essentially cancelled for a short time, and that would mean less ring expansion force and higher MOFT between the ring face and cylinder wall. As the piston moves downward, then the ring starts sealing better between the ring face and wall.
Anyway ... good find on that last paper ... will digest in more detail later, and maybe comment more on your post #96 if I find something. In the middle of doing my taxes, and looks like I owe $$.
Sucks for the taxes. Last year, with the $4,502 plug-in tax credit plus the car sales tax deduction for my 2021 Prius Prime along with the mortgage deductions resulted in a huge refund. I got a refund, too, this year but a much smaller one.
Wanted to comment on this. Even if there is that much combustion pressure behind the top ring making it expand with that much force, it's still possible for the piston to tilt enough (if the piston skirt to wall clearance is excessive) for the area of the ring pack to rub on the cylinder. it also depends a lot on the piston design. That means the ring(s) would have to be totally flush with the piston surface at that point.
I think the part you're missing is that even when there is a large force pushing the top compression ring outward from combustion gas pressure, the ring pack can still float in the piston grooves. If the lateral force on the piston due to the stroke acceleration is great enough, and the piston to wall clearance is also great enough, the piston can still move laterally in the thrust direction regardless of how hard the rings are pushing against the cylinder wall. The rings don't "lock on" to the ring grooves in any case and prevent the piston from rocking back and forth in the thrust directions.
If you watched the video I posted in post #90, that is explained. Also the photo of the piston in post #90, along with the pistons shown in the video at time 7:05, you can see the area above the top ring has wear, so it obviously had to be rocking enough to contact the cylinder wall, and the top ring had to be completely pushed flush with the piston surface in order for that wear to happen.
As that video also explained, it's the piston skirt to cylinder wall clearance and the exact design of the piston that are the main factors that control piston tilting in the bore. If the clearance is large, and/or the piston skirt is very short (typically used in high performance engines), then the piston could certainly tilt enough to rub the top ring area on the cylinder wall. Look at this piston, the skirt is very short, and therefore if the piston to bore clearance isn't pretty tight, it's going to want to tilt much easier than a piston with a much longer shirt. Piston and ring design is truly engineering intensive, as well as an art.
Here are a couple of other examples of what I'm talking about. The pressure behind the top ring didn't ensure the area above the top ring didn't tilt and rub on the wall due to the lateral force applied by the connective rod. The rings still fully float in their grooves regardless of how much combustion pressure is behind them.
I agree ... I'm always up for a civil technical discussion, and it does also become a learning experience because we have to go dig up all the technical references to debate viewpoints.This discussion veered quite off from the effect of the base-oil viscosity and VII on the lubrication of the valvetrain and piston rings, but we've learned quite a bit. I say you should believe that the shear rates in the valvetrain and at the rings can be as quite high as the Shell paper I posted earlier reported.
I do believe the shear rates in the Shell paper when they assume super small MOFT. But I also understand that their peak shear rate numbers are most likely not going to be seen very often if at all on a normally driven street car that lives between 2,000 and 3,000 RPM 98% of the time. For the track guys, racers and heavy towers (much hotter oil temps), obviously they need to bump the viscosity up to get that needed higher HTFS, even if the higher viscosity grade oil has a larger HTHS to HTFS reduction due to VIIs and high shear rates.
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OVERKILL
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Great discussion indeed gents, enjoyed reading it thus far.
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