Okay, stupid question. After looking at this article (whew) am I correct that in a manual transmission car one is better off in terms of bearing wear (usually the lead content) to keep the RPMs above a certain range, say 2000-2300- RPMs (obviously engine specific) but to keep the RRMs above the rate where the engine would lug. It appears that lugging causes bearing wear which could be reduced by keepping the transmission in a lower gear to keep the RPMS up. Most people want to get to the higher gear ASAP for gas mileage or whatever reason but a higher RPM may actually be better for bearing wear. Is this a correct assessment? And of course, downshift when necessary in lieu of lugging.
Determining Oil Viscosity by Lasche-McKee
- Thread starter Spector
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Spector,
You are indeed correct that lugging the engine - especially under heavy load - can increase bearing wear. Hard acceleration from low rpms is a particularly bad thing to do. Oil pressure is a function of engine rpms and you have fairly low oil pressure at 2000 rpms. However, keep in mind also that every rpm is a wear cycle, so there is a tradeoff here. For example, piston ring and cylinder wear correlates well to average piston speeds, which are in turn related to the stroke of the engine and the rpm range in which it runs. For this reason, high rpm engines are often "oversquare" designs, with large bores and short strokes to minimize piston speeds. If you consider the change in momentum that occurs at the top and bottom of each stroke, this makes perfect sense.
Some of the slowest wearing engines are medium to large diesels that run at low rpms. These engine also spec 15w-40 oils instead of 5w-20/5w-30 grades to compensate for the low rpm operation and provide higher oil pressure and oil film thickness.
For a given situation, oil film thickness is a function of the load between parts, the viscosity of the oil and the relative velocity between the opposing surfaces. A lower relative velocity decreases the oil film thickness. This is true for cam lobe/lifter oil films as well as for main/rod bearing films.
[ April 16, 2003, 07:11 PM: Message edited by: TooSlick ]
You are indeed correct that lugging the engine - especially under heavy load - can increase bearing wear. Hard acceleration from low rpms is a particularly bad thing to do. Oil pressure is a function of engine rpms and you have fairly low oil pressure at 2000 rpms. However, keep in mind also that every rpm is a wear cycle, so there is a tradeoff here. For example, piston ring and cylinder wear correlates well to average piston speeds, which are in turn related to the stroke of the engine and the rpm range in which it runs. For this reason, high rpm engines are often "oversquare" designs, with large bores and short strokes to minimize piston speeds. If you consider the change in momentum that occurs at the top and bottom of each stroke, this makes perfect sense.
Some of the slowest wearing engines are medium to large diesels that run at low rpms. These engine also spec 15w-40 oils instead of 5w-20/5w-30 grades to compensate for the low rpm operation and provide higher oil pressure and oil film thickness.
For a given situation, oil film thickness is a function of the load between parts, the viscosity of the oil and the relative velocity between the opposing surfaces. A lower relative velocity decreases the oil film thickness. This is true for cam lobe/lifter oil films as well as for main/rod bearing films.
[ April 16, 2003, 07:11 PM: Message edited by: TooSlick ]
MolaKule
Staff member
Addendum:
The Note on page 2 should read:
Note: The algorithm presented here does overstate the viscosity for Synthetic Fluids by 12%, but does get one into the ballpark. This algorithm was developed in the days when all lubricants were petroleum.
Also, most engineers today use a modified version of the equations presented in the paper to account for certain subtleties in the particular engine design, and the modified algorithms are usually programmed into and integrated with a computer code that includes Finite Element Analysis (FEA), CAD, etc.
[ April 16, 2003, 07:52 PM: Message edited by: MolaKule ]
The Note on page 2 should read:
Note: The algorithm presented here does overstate the viscosity for Synthetic Fluids by 12%, but does get one into the ballpark. This algorithm was developed in the days when all lubricants were petroleum.
Also, most engineers today use a modified version of the equations presented in the paper to account for certain subtleties in the particular engine design, and the modified algorithms are usually programmed into and integrated with a computer code that includes Finite Element Analysis (FEA), CAD, etc.
[ April 16, 2003, 07:52 PM: Message edited by: MolaKule ]
MolaKule
Staff member
A number of people have requested the tutorial on bearings
and the method for determining oil viscosity required for bearings.
BobIs has been gracious to upload the MsWord document to this site:
http://users.ms11.net/~pic/doc/Lasche_McKee.doc
If you have any questions or comments, you can post them
here or send me a PM.
Mola
[ April 16, 2003, 12:34 PM: Message edited by: MolaKule ]
and the method for determining oil viscosity required for bearings.
BobIs has been gracious to upload the MsWord document to this site:
http://users.ms11.net/~pic/doc/Lasche_McKee.doc
If you have any questions or comments, you can post them
here or send me a PM.
Mola
[ April 16, 2003, 12:34 PM: Message edited by: MolaKule ]
MolaKule
Staff member
Here is the Matlab Code for Same:
% Lasche-McKee Bearing/Viscosity Algorithm
% by MolaKule
% Example is for Chevy 350 V8 Engine
% Tb is max bearing temp
Tb = 230
% To is Oil temp
To = 190
delT = Tb - To
% C is Diametral Clearance
C = 0.001925
% Z is viscosity in centipoise
Z = 1
% k is constant read from L/D graph
k = 0.006
% K is constant from bearing type,
% heat dissipation condition, etc.,
% K ranges from 31 to 55
K = 35
% N is RPM
N = 3000
% L is Length (or width) of Bearing in in.
L = 1.21875
% D is diameter of Journal in in. from Spec. sheets
D = 2.448
% W is Radial bearing load in lbs.
W = 1200
% p is net pressure = W/LD
p = W/(L*D)
% Next determine heat dissipated in Bearing
Hd = ((delT + 33)^2/K)*L*D
% determine each term of friction coefficient
% f = k + Z*4.73e-8*(N/p)*(D/C)
% FSTZ = second term of friction equation
FSTZ = 4.73e-8*(N/p)*(D/C)
% Next determine Heat generated in bearing
% Hg = ((f*W*pi*D*N)/12)
% Determine WPDN, WPDN = (W*pi*D*N)/12
WPDN = (W*pi*D*N)/12
% HgFT1 is First term of friction equation X Hg, k*WPDN
% Since you have two terms of the Hg equation.
% HgFT2 is Second term of friction equation X Hg, FSTZ*WPDN
HgFT1 = k*WPDN
HgFT2 = FSTZ*WPDN
Zprime = HgFT1 - Hd
% Determine Z in centipoise, cP
Z = Zprime/HgFT2
% Use Chart or Formula to Determine Kinematic Viscosity
% Lasche-McKee Bearing/Viscosity Algorithm
% by MolaKule
% Example is for Chevy 350 V8 Engine
% Tb is max bearing temp
Tb = 230
% To is Oil temp
To = 190
delT = Tb - To
% C is Diametral Clearance
C = 0.001925
% Z is viscosity in centipoise
Z = 1
% k is constant read from L/D graph
k = 0.006
% K is constant from bearing type,
% heat dissipation condition, etc.,
% K ranges from 31 to 55
K = 35
% N is RPM
N = 3000
% L is Length (or width) of Bearing in in.
L = 1.21875
% D is diameter of Journal in in. from Spec. sheets
D = 2.448
% W is Radial bearing load in lbs.
W = 1200
% p is net pressure = W/LD
p = W/(L*D)
% Next determine heat dissipated in Bearing
Hd = ((delT + 33)^2/K)*L*D
% determine each term of friction coefficient
% f = k + Z*4.73e-8*(N/p)*(D/C)
% FSTZ = second term of friction equation
FSTZ = 4.73e-8*(N/p)*(D/C)
% Next determine Heat generated in bearing
% Hg = ((f*W*pi*D*N)/12)
% Determine WPDN, WPDN = (W*pi*D*N)/12
WPDN = (W*pi*D*N)/12
% HgFT1 is First term of friction equation X Hg, k*WPDN
% Since you have two terms of the Hg equation.
% HgFT2 is Second term of friction equation X Hg, FSTZ*WPDN
HgFT1 = k*WPDN
HgFT2 = FSTZ*WPDN
Zprime = HgFT1 - Hd
% Determine Z in centipoise, cP
Z = Zprime/HgFT2
% Use Chart or Formula to Determine Kinematic Viscosity
Too Slick said: Some of the slowest wearing engines are medium to large diesels that run at low rpms. These engine also spec 15w-40 oils instead of 5w-20/5w-30 grades to compensate for the low rpm operation and provide higher oil pressure and oil film thickness.
For a given situation, oil film thickness is a function of the load between parts, the viscosity of the oil and the relative velocity between the opposing surfaces. A lower relative velocity decreases the oil film thickness. This is true for cam lobe/lifter oil films as well as for main/rod bearing films. This said is this why when they design a engine the manufacture sets the viscosity of the oil to be recommended.As the bearings wear the clearance increaces would the oil pressure show at any operating RPM that a thicker viscosity oil is needed or is the specs designed for the life of the engine.
For a given situation, oil film thickness is a function of the load between parts, the viscosity of the oil and the relative velocity between the opposing surfaces. A lower relative velocity decreases the oil film thickness. This is true for cam lobe/lifter oil films as well as for main/rod bearing films. This said is this why when they design a engine the manufacture sets the viscosity of the oil to be recommended.As the bearings wear the clearance increaces would the oil pressure show at any operating RPM that a thicker viscosity oil is needed or is the specs designed for the life of the engine.
MolaKule
Staff member
I think he was referring to the condition of low RPM AND high loads, such as lugging. One would certainly want more oil flow during high loads to increase film thickness and keep the oil cooler, which go hand in hand.
As far as the mass flow of oil, I stated elsewhere this was in error. This example was for a SAE 20 weight oil, a 1.5" dia. journal in a bearing turning at 30 rev/s (1,800 rpm), the load was 500 pounds, and the length of the bearing was 1.5".
I believe the author mixed units here and never gave his computer (code) listings or references as to the exact equations used. The answer I calculated for mass flow was 150 mL/s in metric units, for a temperature rise in the bearing's oil of 28 F.
In most bearing oil temp rise calculations, the temp rise is for 20 C minimum.
As far as the mass flow of oil, I stated elsewhere this was in error. This example was for a SAE 20 weight oil, a 1.5" dia. journal in a bearing turning at 30 rev/s (1,800 rpm), the load was 500 pounds, and the length of the bearing was 1.5".
I believe the author mixed units here and never gave his computer (code) listings or references as to the exact equations used. The answer I calculated for mass flow was 150 mL/s in metric units, for a temperature rise in the bearing's oil of 28 F.
In most bearing oil temp rise calculations, the temp rise is for 20 C minimum.
First, nice paper MolaKule! It makes a complex topic about as clear as any explanation I have seen. I didn't read it too closely, but noticed that the flow you quoted looked a wee bit high.
"Shigley and Mischke, in "Mechanical Engineering Design," show that the oil flow required to maintain a 28 0F differential between oil-in temp and oil-out temp of a bearing of 0.00175" clearance, is Q = 0.252 m3/s. "
TooSlick, good response except you give the inmpression that lower oil pressure at lower RPMs is a problem. Not necessarily so. At lower RPMs a journal bearing throw off less oil, so needs less oil. All the pump pressure does is get the oil to the bearing. As long as the bearing gets an adequate supply, it doesn't matter what the pressure was. Oil pump pressure doesn't have any direct effect on how well a journal bearing works as long as enough oil gets to the bearing. I think you understand that, but your post wasn't clear on that so might have mislead the casual reader.
does this mean that the ford 5.4l engines would do better (wear wise) w/ a 30 or 40 weight oil instead of the 20 weight???
Yes, but....
At lower rpm there is more time for the void on the unloaded side of the bearing to get filled with oil, so less pressure is needed to fill the void. As long as the void gets filled, or nearly filled, it makes no difference what oil pressure you have.
The old rule of thumb of 10 psi/1000 rpm was actually pretty good in that respect.
It shouldn't make much difference uless you break through the oil film, and you shouldn't break through the oil film if all is correct.
What does work against you with 20 weight is that there is less reserve if something isn't quite right.
MolaKule
Staff member
XS,
True, but you are only considering one of the two sufficient conditions for bearing longevity - a minimum required oil film thickness for hydrodynamic or mixed-film lubrication.
The other side of the equation is bearing cooling. You need a pressurized oil to create a flow of fluid to cool the bearing, hence the 150 mL/s of oil flow for the conditions cited (not really that much oil flow, BTW). Any bearing that is buried in a pool of oil will suck in it's own oil needs, but it will not cool as efficiently as a pressurized oil system.
True, but you are only considering one of the two sufficient conditions for bearing longevity - a minimum required oil film thickness for hydrodynamic or mixed-film lubrication.
The other side of the equation is bearing cooling. You need a pressurized oil to create a flow of fluid to cool the bearing, hence the 150 mL/s of oil flow for the conditions cited (not really that much oil flow, BTW). Any bearing that is buried in a pool of oil will suck in it's own oil needs, but it will not cool as efficiently as a pressurized oil system.
You have made me think, I shall be wary of you in the future
I was overlooking cooling flow, but.... cooling requirements also go down almost linearly with rpm, so less pressure is still required to cool the bearing at lower rpm.
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