LS1 Oil Pump Pressure Relief Study
LS1 Oil Pump Pressure Relief Study I am in the process of a head/cam install that includes a new oil pump. I knew some folks were shimming their pressure relief springs to increase system oil pressure. From my small/big block Chevy days, I remember some replacement pumps shipping with a different pressure relief spring for certain applications. Researching the internet on shimming the LS1 oil pump produced many recommendations but nothing seemed to be based on science, they were “seat-of-the-pants” decisions. Being a Mechanical Engineer, I decided to tackle the subject myself. However my speciallity is HVAC design NOT fluid flow (at least not hydraulic fluid-type applications).
The following is a summary of my research. Please feel free to comment, criticize, question, or expose any fallacies. Where possible, I listed sources for formulas and attempted to point out critical assumptions. A copy of this study is available at: http://www.hodge-rulez.com/ls6_pump.htm.
The following study has not been proven, major news outlets have not bashed it, the Republi-Crat Party has not deemed it vital to national security or imprisoned/tortured on behalf of, but Alfred E. Neuman HAS approved it for publication.
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LS1 Oil Pump Pressure Relief Study
Objective: Calculate the theoretical oil pressure the LS1 oil pump should maintain. Compare results to actual oil pressure (in 2000 Corvette) & printed GM specifications. Use these results to develop an equation to calculate the required relief spring shim to produce a desired system oil pressure.
Background: The so-called “LS6” oil pump, GM p/n 12586665, was disassembled and the physical dimensions of the relief piston, spring, and port are:
Figure 1 – Relief Valve Piston/Spring Assembly
D = Piston Diameter = 0.570 inch
d = Spring wire diameter = 0.0535 inch
Dmean = mean diameter of spring coils = 0.3365 inch
xv = piston displacement to pressure relief port = 0.135 inch
x0 = precompressed spring length = 0.227 inch
The force a spring produces is defined by the relationship:
Where k is the spring constant and the displacement is the change in spring length.
The piston relief spring constant, ks, is defined by (from “Mechanical Engineering Design, Fifth Edition” by Shigley & Mischke):
where d = diameter of the spring wire
G = Modulus of Rigidity
D = mean diameter of the spring coils
Na = number of active coils = number of total coils, Nt, minus 2 for a spring with squared and ground ends (Na = Nt – 2)
Assumption #1 – Assume the spring is made of chrome vanadium wire. This wire material has a maximum design temperature rating of 425°F. (from http://www.engineersedge.com/spring_general.htm)
G = 11.5 x 106 psi (psi = lbf/in2)
d = .0535 inch
D = 0.3365 inch
Na = 16 – 2 = 14
Substituting the above into equation (3) and solving for ks yields:
The pressure relief piston design is similar to a direct acting relief valve in a hydraulic spool valve. If you assume a static condition, the sum of the forces must equal zero. The forces on the piston are:
Figure 2 – Forces Acting on Pressure Relief Piston
The equation defining the above is (from “Hydraulic Component Design and Selection” by Fitch & Hong):
Where Ps = system pressure
As = piston area (piston diameter = 0.570 inch) = .2552 inch2
ks = piston spring constant
x0 = precompressed spring length = 0.227 inch
xv = valve opening width (piston displacement to pressure relief port) = 0.135 inch
Fsf = steady state flow force of the motor oil through the pressure relief port
Fr = Coulomb friction force
Piston area, As, = pD2/4 = p(0.570 inch)2/4 = 0.255 inch2
Assumption #3 – assume the force caused by the friction between the piston and the pressure port is much lower than the other forces, i.e., Fr @ 0.
If Fr = 0, equation (2) becomes:
The steady state flow force (Fsf) in equation (2) is defined by:
Where Cf = flow force area gradient
The flow force area gradient (axial or perpendicular to the piston) is defined by:
Where D = piston diameter = 0.570 inch
Cd = flow discharge coefficient
Cv = flow velocity coefficient
f = flow jet angle
Combining equations (4) and (5) yields:
from http://www.machinedesign.com/asp/vie...MDSite&catId=2,
Assumption #4, 5, & 6:
Cd = 0.7
Cv = 1
f = 69°
Solving equation (3) for Ps yields:
Substituting equation (6) into (7) yields:
Combining terms and solving equation (8) for Ps yields:
Summary of the terms in equation (9) are:
ks = 21.885 lbf/in2 (psi)
x0 = 0.227 inch
xv = 0.135 inch
As = 0.2552 inch2
D = 0.570 inch
Cd = 0.7
Cv = 1
f = 69°
Substituting the above into equation (9) yields:
The calculated system pressure, Ps, is close to the design oil pump pressure I have seen on the internet (somewhere I saw 60 psi as the design pressure; sorry I cannot list the source). My 2000 Corvette has a system pressure of about 58/59 psi at 5,000 rpm. Since the calculated and actual pressure is very close (less than 1% error), I must conclude equation (9) is accurate for the purposes of further discussion.
The following is a summary of my research. Please feel free to comment, criticize, question, or expose any fallacies. Where possible, I listed sources for formulas and attempted to point out critical assumptions. A copy of this study is available at: http://www.hodge-rulez.com/ls6_pump.htm.
The following study has not been proven, major news outlets have not bashed it, the Republi-Crat Party has not deemed it vital to national security or imprisoned/tortured on behalf of, but Alfred E. Neuman HAS approved it for publication.
__________________________________________________ _______
LS1 Oil Pump Pressure Relief Study
Objective: Calculate the theoretical oil pressure the LS1 oil pump should maintain. Compare results to actual oil pressure (in 2000 Corvette) & printed GM specifications. Use these results to develop an equation to calculate the required relief spring shim to produce a desired system oil pressure.
Background: The so-called “LS6” oil pump, GM p/n 12586665, was disassembled and the physical dimensions of the relief piston, spring, and port are:
Figure 1 – Relief Valve Piston/Spring Assembly
(For illustrative purposes only; not to scale)
D = Piston Diameter = 0.570 inch
d = Spring wire diameter = 0.0535 inch
Dmean = mean diameter of spring coils = 0.3365 inch
xv = piston displacement to pressure relief port = 0.135 inch
x0 = precompressed spring length = 0.227 inch
The force a spring produces is defined by the relationship:
Force = k x displacement
Where k is the spring constant and the displacement is the change in spring length.
The piston relief spring constant, ks, is defined by (from “Mechanical Engineering Design, Fifth Edition” by Shigley & Mischke):
ks = (d4 * G)/(8*D3*Na) (1)
where d = diameter of the spring wire
G = Modulus of Rigidity
D = mean diameter of the spring coils
Na = number of active coils = number of total coils, Nt, minus 2 for a spring with squared and ground ends (Na = Nt – 2)
Assumption #1 – Assume the spring is made of chrome vanadium wire. This wire material has a maximum design temperature rating of 425°F. (from http://www.engineersedge.com/spring_general.htm)
G = 11.5 x 106 psi (psi = lbf/in2)
d = .0535 inch
D = 0.3365 inch
Na = 16 – 2 = 14
Substituting the above into equation (3) and solving for ks yields:
ks = 21.885 lbf/in2 (psi)
The pressure relief piston design is similar to a direct acting relief valve in a hydraulic spool valve. If you assume a static condition, the sum of the forces must equal zero. The forces on the piston are:
- The force of the system oil pressure attempting to open the valve
- The force of the piston spring preload
- The force of the piston spring deflection to the pressure relief port
- The force caused by the flow of oil across the piston
- The friction of the piston to the oil pump piston bore
Figure 2 – Forces Acting on Pressure Relief Piston
The equation defining the above is (from “Hydraulic Component Design and Selection” by Fitch & Hong):
PsAs = ks(x0 + xv) +Fsf - Fr (2)
Where Ps = system pressure
As = piston area (piston diameter = 0.570 inch) = .2552 inch2
ks = piston spring constant
x0 = precompressed spring length = 0.227 inch
xv = valve opening width (piston displacement to pressure relief port) = 0.135 inch
Fsf = steady state flow force of the motor oil through the pressure relief port
Fr = Coulomb friction force
Piston area, As, = pD2/4 = p(0.570 inch)2/4 = 0.255 inch2
Assumption #3 – assume the force caused by the friction between the piston and the pressure port is much lower than the other forces, i.e., Fr @ 0.
If Fr = 0, equation (2) becomes:
PsAs = ks(x0 + xv) +Fsf (3)
The steady state flow force (Fsf) in equation (2) is defined by:
Fsf = CfPsxv (4)
Where Cf = flow force area gradient
The flow force area gradient (axial or perpendicular to the piston) is defined by:
Cf = 2pDCdCvcosf (5)
Where D = piston diameter = 0.570 inch
Cd = flow discharge coefficient
Cv = flow velocity coefficient
f = flow jet angle
Combining equations (4) and (5) yields:
Fsf = 2pDCdCvPsxvcosf (6)
from http://www.machinedesign.com/asp/vie...MDSite&catId=2,
Assumption #4, 5, & 6:
Cd = 0.7
Cv = 1
f = 69°
Solving equation (3) for Ps yields:
Ps = {ks(x0 + xv) +Fsf}/As (7)
Substituting equation (6) into (7) yields:
Ps = {ks(x0 + xv) + 2pDCdCvPsxvcosf}/As (8)
Combining terms and solving equation (8) for Ps yields:
Ps = ks(x0 + xv) / (As + 2pDCdCvxvcosf) (9)
Summary of the terms in equation (9) are:
ks = 21.885 lbf/in2 (psi)
x0 = 0.227 inch
xv = 0.135 inch
As = 0.2552 inch2
D = 0.570 inch
Cd = 0.7
Cv = 1
f = 69°
Substituting the above into equation (9) yields:
Ps = 59.2 psi
The calculated system pressure, Ps, is close to the design oil pump pressure I have seen on the internet (somewhere I saw 60 psi as the design pressure; sorry I cannot list the source). My 2000 Corvette has a system pressure of about 58/59 psi at 5,000 rpm. Since the calculated and actual pressure is very close (less than 1% error), I must conclude equation (9) is accurate for the purposes of further discussion.
Study Continued; Part II Back when I use to play with Chevelles and Camaros, I used a rule-of-thumb of 10 psi oil pressure for each 1,000 rpms of engine operation. This rule-of-thumb is still valid today. What I want to know is how much to shim the oil pump spring to produce a desired system pressure. Now for a more “scientific” calculation of an oil pump spring shim:
Equation (2) defines the spring force as:
If a shim is added to the piston spring to increase preload, Xshim, equation (10) becomes:
Substitute equation (11) into equation (9) yields:
Solving equation (12) for Xshim yields:
Substituting all the known terms into equation (13) reduces to:
Equation (14) gives the required shim thickness if the system pressure is entered in units of psi. This equation assumes the system pressure is 59 psi. If you input a pressure less than 59 psi, Xshim will be negative.
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Example:
The design max rpm (redline) after a head/cam swap 7,000 (assuming the valve train will support this rpm). The desired system oil pressure is 70 psi. Plugging this into equation (14) yields:
xshim = 0.00612(70) – 0.362
The oil pump relief spring needs to be shimmed 0.066 inch to produce a 70 psi system oil pressure.
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I have seen recommendations to shim the oil pump relief spring anywhere from 0.060 to 0.120 inches. Seems this produces 70 to 80 psi oil pressure. If you put 80 psi in equation (14), you get a shim thickness of 0.128 inch. Again, this seems to match the recommendations I have seen on the Internet.
I plan to shim my oil pump 0.060 inches. Once I get the engine in the car, I will report the system oil pressure.
Equation (2) defines the spring force as:
Force = ks(x0 + xv) (10)
If a shim is added to the piston spring to increase preload, Xshim, equation (10) becomes:
Force = ks(x0 + xv + xshim) (11)
Ps = ks(x0 + xv + xshim) / (As + 2pDCdCvxvcosf) (12)
Solving equation (12) for Xshim yields:
xshim = PsAs/ks - 2pDCdCvxvcosf/ks - x0 - xv (13)
Substituting all the known terms into equation (13) reduces to:
xshim = 0.00612Ps – 0.362 (14)
Equation (14) gives the required shim thickness if the system pressure is entered in units of psi. This equation assumes the system pressure is 59 psi. If you input a pressure less than 59 psi, Xshim will be negative.
_______________________________________________
Example:
The design max rpm (redline) after a head/cam swap 7,000 (assuming the valve train will support this rpm). The desired system oil pressure is 70 psi. Plugging this into equation (14) yields:
xshim = 0.00612(70) – 0.362
xshim = 0.066 inch
The oil pump relief spring needs to be shimmed 0.066 inch to produce a 70 psi system oil pressure.
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I have seen recommendations to shim the oil pump relief spring anywhere from 0.060 to 0.120 inches. Seems this produces 70 to 80 psi oil pressure. If you put 80 psi in equation (14), you get a shim thickness of 0.128 inch. Again, this seems to match the recommendations I have seen on the Internet.
I plan to shim my oil pump 0.060 inches. Once I get the engine in the car, I will report the system oil pressure.
Seems to me that when the oil pressure is below the design pressure (idle), shimming the spring would have no effect because the bypass is completely closed anyway. In other words, shimming the spring would not increase your oil pressure at idle. True?
Originally Posted by P Mack
Seems to me that when the oil pressure is below the design pressure (idle), shimming the spring would have no effect because the bypass is completely closed anyway. In other words, shimming the spring would not increase your oil pressure at idle. True?
Looking back at the vehicles I've owned, the one that was the most fun to drive was a '66 El Camino with a really hot 327. This engine had under 20 psi oil pressure at idle. At 3,000 rpm it had 60 psi. I ran that engine HARD for years. Never let me down.
TECH Regular
Joined: Oct 2005
Posts: 430
Likes: 0
FWIW, Pontiac V-8s were notorious for low idle oil pressure. In fact, I'm pretty sure none were factory-equipped with actual gauges for this reason. I've seen low mileage Ponchos, e.g. 1970 400 c.i. GTO, that ran down the road at 50 psi but showed barely 5 psi hot idle pressure.. 
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Originally Posted by MadBill
FWIW, Pontiac V-8s were notorious for low idle oil pressure. In fact, I'm pretty sure none were factory-equipped with actual gauges for this reason. I've seen low mileage Ponchos, e.g. 1970 400 c.i. GTO, that ran down the road at 50 psi but showed barely 5 psi hot idle pressure..
Originally Posted by gun5l1ng3r
^Was it just the design of the pump?
Where the pump only flowed enough to create 10-15 PSI at idle, but as RPM increased, the raise in PSI increased fast.
Where the pump only flowed enough to create 10-15 PSI at idle, but as RPM increased, the raise in PSI increased fast.
Hodge,
I agree with your statements in general. However, I would advocate porting the oil pump inlet and outlet without a shim since you already have excellent oil pressure. Take a look at a recent post of mine where I touch on some of these issues https://ls1tech.com/forums/showpost....87&postcount=5.
Steve
I agree with your statements in general. However, I would advocate porting the oil pump inlet and outlet without a shim since you already have excellent oil pressure. Take a look at a recent post of mine where I touch on some of these issues https://ls1tech.com/forums/showpost....87&postcount=5.
Steve
Teching In
Joined: Dec 2005
Posts: 15
Likes: 0
From: australia
Too many people get hung up on oil pressure.Oil FLOW is really what counts.
How many times have i read of people changing to a higher weight oil and being impressed by the higher indicated oil pressure.
WRONG.This just means that the flow has been reduced and flow is more important than pressure as long as the pressure is within the manufactures limits.10psi/1000 revs is a pretty good yardstick.
Ciao
How many times have i read of people changing to a higher weight oil and being impressed by the higher indicated oil pressure.
WRONG.This just means that the flow has been reduced and flow is more important than pressure as long as the pressure is within the manufactures limits.10psi/1000 revs is a pretty good yardstick.
Ciao
Originally Posted by Steve Bryant
Hodge,
I agree with your statements in general. However, I would advocate porting the oil pump inlet and outlet without a shim since you already have excellent oil pressure. Take a look at a recent post of mine where I touch on some of these issues https://ls1tech.com/forums/showpost....87&postcount=5.
Steve
I agree with your statements in general. However, I would advocate porting the oil pump inlet and outlet without a shim since you already have excellent oil pressure. Take a look at a recent post of mine where I touch on some of these issues https://ls1tech.com/forums/showpost....87&postcount=5.
Steve
Hodge,
The pressure regulator regulates by returning excess oil to the intake side of the gerator pump when the spring loaded piston is displaced to the point that the slit/orifice in the side of the cylinder wall is opened by the oil pressure. If the pump is putting out sufficient volume with the oil at a given viscosity and the oil receiving capacity of the engine is maximized (collective sum of oil clearance flow plus return to the sump plus bypassed oil via the pressure regulator), the pressure will increase. Let's say that the pressure regulator can bypass 6 GPM at a given stroke position of the piston, but it needs to bypass 8 to regulate @ 60 PSI. The pressure will increase beyond 60 PSI. If the bearing clearances, lifter clearances, etc. were greater, this pressure increase would not occur. Also, ask yourself if you have 58/59 PSI now, what would you gain by additional pressure. Would the oiling be improved? Probably not. Would the parasitic loss attributable to the oil pump increase? Probably yes.
Just my two cents.
Steve
The pressure regulator regulates by returning excess oil to the intake side of the gerator pump when the spring loaded piston is displaced to the point that the slit/orifice in the side of the cylinder wall is opened by the oil pressure. If the pump is putting out sufficient volume with the oil at a given viscosity and the oil receiving capacity of the engine is maximized (collective sum of oil clearance flow plus return to the sump plus bypassed oil via the pressure regulator), the pressure will increase. Let's say that the pressure regulator can bypass 6 GPM at a given stroke position of the piston, but it needs to bypass 8 to regulate @ 60 PSI. The pressure will increase beyond 60 PSI. If the bearing clearances, lifter clearances, etc. were greater, this pressure increase would not occur. Also, ask yourself if you have 58/59 PSI now, what would you gain by additional pressure. Would the oiling be improved? Probably not. Would the parasitic loss attributable to the oil pump increase? Probably yes.
Just my two cents.
Steve
Originally Posted by lucky
Too many people get hung up on oil pressure.Oil FLOW is really what counts.
How many times have i read of people changing to a higher weight oil and being impressed by the higher indicated oil pressure.
WRONG.This just means that the flow has been reduced and flow is more important than pressure as long as the pressure is within the manufactures limits.10psi/1000 revs is a pretty good yardstick.
Ciao
How many times have i read of people changing to a higher weight oil and being impressed by the higher indicated oil pressure.
WRONG.This just means that the flow has been reduced and flow is more important than pressure as long as the pressure is within the manufactures limits.10psi/1000 revs is a pretty good yardstick.
Ciao
Originally Posted by Steve Bryant
Hodge,
The pressure regulator regulates by returning excess oil to the intake side of the gerator pump when the spring loaded piston is displaced to the point that the slit/orifice in the side of the cylinder wall is opened by the oil pressure. If the pump is putting out sufficient volume with the oil at a given viscosity and the oil receiving capacity of the engine is maximized (collective sum of oil clearance flow plus return to the sump plus bypassed oil via the pressure regulator), the pressure will increase. Let's say that the pressure regulator can bypass 6 GPM at a given stroke position of the piston, but it needs to bypass 8 to regulate @ 60 PSI. The pressure will increase beyond 60 PSI. If the bearing clearances, lifter clearances, etc. were greater, this pressure increase would not occur. Also, ask yourself if you have 58/59 PSI now, what would you gain by additional pressure. Would the oiling be improved? Probably not. Would the parasitic loss attributable to the oil pump increase? Probably yes.
Just my two cents.
Steve
The pressure regulator regulates by returning excess oil to the intake side of the gerator pump when the spring loaded piston is displaced to the point that the slit/orifice in the side of the cylinder wall is opened by the oil pressure. If the pump is putting out sufficient volume with the oil at a given viscosity and the oil receiving capacity of the engine is maximized (collective sum of oil clearance flow plus return to the sump plus bypassed oil via the pressure regulator), the pressure will increase. Let's say that the pressure regulator can bypass 6 GPM at a given stroke position of the piston, but it needs to bypass 8 to regulate @ 60 PSI. The pressure will increase beyond 60 PSI. If the bearing clearances, lifter clearances, etc. were greater, this pressure increase would not occur. Also, ask yourself if you have 58/59 PSI now, what would you gain by additional pressure. Would the oiling be improved? Probably not. Would the parasitic loss attributable to the oil pump increase? Probably yes.
Just my two cents.
Steve
TECH Enthusiast
Joined: Jan 2004
Posts: 633
Likes: 1
From: Detroit,MI
The biggest point that was improved upon when Melling released their LS1 pump was the increase in oil volume, pressure is a byproduct of increased volume.
They accomplished this by using thicker components in the pump that will physically move more oil, and they also have standard and high pressure springs that are available.
Good research above.
They accomplished this by using thicker components in the pump that will physically move more oil, and they also have standard and high pressure springs that are available.
Good research above.
Teching In
Joined: Jan 2006
Posts: 18
Likes: 0
From: ShenZhen, China
Great work- Can I interest you in another project to do some airflow studies for my skunkworks sheetmetal intake? Skunkworks because if I do it on my own, it will probably stink. 
Not kidding about the work; I like seeing people put things together like that and presnet the facts or as close as you can get. Excellent post and very cool to the hobby....
PM me if you want to develop that intake.
Not kidding about the work; I like seeing people put things together like that and presnet the facts or as close as you can get. Excellent post and very cool to the hobby....
PM me if you want to develop that intake.
Originally Posted by phaseshift
Great work- Can I interest you in another project to do some airflow studies for my skunkworks sheetmetal intake? Skunkworks because if I do it on my own, it will probably stink.
Not kidding about the work; I like seeing people put things together like that and presnet the facts or as close as you can get. Excellent post and very cool to the hobby....
PM me if you want to develop that intake.
Not kidding about the work; I like seeing people put things together like that and presnet the facts or as close as you can get. Excellent post and very cool to the hobby....
PM me if you want to develop that intake.
source code for the ECM
design requirements for heads
camshaft design
oil pump design
communication protocol between the various on-board computers
Too bad GM is a for-profit company...
Sure would make some things easier. However, I do believe "necessity is the mother of invention." It is amazing what you are capable of when you must solve a problem.
One of my core strenghts at work is "indoor air moving" (blower design). That is a far cry from sheet metal intake and air distrubition in an engine. I'll pass
