Don Terrill’s Speed-Talk

Does anyone have a practical formula for minimum required exhaust valves seated load as a function of component weights, camshaft profile, rpm, valve diameter, and exhaust back pressure?

ptuomov wrote: ↑Sun Oct 27, 2019 4:50 pm Does anyone have a practical formula for minimum required exhaust valves seated load as a function of component weights, camshaft profile, rpm, valve diameter, and exhaust back pressure?

Years ago I wrote a spreadsheet to compute valve spring force needed to prevent valve float, which usually caused valve bounce upon closing.
It was based on valve ensemble weight, spring mass, valve acceleration at various engine rpm, valve lift, installed height, spring rate, coil bind and spring natural frequency.

It did not include, or provide for, inlet or exhaust air pressures.

ptuomov wrote: ↑Sun Oct 27, 2019 4:50 pm Does anyone have a practical formula for minimum required exhaust valves seated load as a function of component weights, camshaft profile, rpm, valve diameter, and exhaust back pressure?

There are a lot of papers on how to estimate.
The first good one is Barken 1954?

Don't let the test mislead you the computation is very complex.

The velocity at contact and the spring in the head disc are two major factors.

The velocity at contact and the spring in the head disc are two major factors.

They are indeed but follow velocity at separation in importance.

I agree that it’s complicated and it’s difficult to measure all the inputs. Let’s make this simpler and more practical. Four valve DOHC engine with direct acting bucket hydraulic lifters. If we have a known good baseline that just barely doesn’t bounce the exhaust valve at a given rpm and boost, what’s the approximate formula to find the new critical exhaust valve seated load if I change:
- reciprocating component weights
- rpm
- velocity (per crankshaft degrees) at which the camshaft closes the valve
- spring resonant frequency (as specified by the spring manufacturer)
- exhaust valve head diameter
- boost and therefore the exhaust back pressure
?

—-

On the exhaust back pressure, I used Vannik’s simulator. The simulated traces reveal that with a cross plane V8 that runs exhaust manifolds the pressure differential at EVC between the exhaust port and the cylinder is completely different between cylinders and the patterns vary with rpm. I graphed those traces for every 500 rpm increment for every cylinder for two boost levels to estimate how much more seated load the engine needs when going from 1.1bar boost to 1.8bar boost.

Example:
Cylinder #5 at 5500 rpm, which is the worst at that rpm. At 1.8bar boost, we're seeing 2.75 bar difference, while at 1.1 bar boost we're seeing 2.1 bar difference.

Now, a simple empirical formula based on the simulations. The additional exhaust valve spring seated load in lbf required because of exhaust pressure waves is approximately 1.26 sqin *14.7 lbf/sqin *2.75 * (1+boost in bar)/(2.8 bar). This is exhaust valve area times ambient pressure times the worst case pressure ratio at 1.8 bar boost scaled to boost you're running.

At 1.8 bar boost, the formula gives 51 lbf. At 1.1 bar boost, the formula gives 38 lbf. Since the car ran fine at 1.1 bar boost and current exhaust spring is set to about 80 lbf seated load, 93 lbf exhaust seated load should work for 1.8 bar boost.

I admit that this isn’t doing it from the first principles, but I think there’s a meaningful area between “from the first principles” and “obviously stupid”.

The problem is how do you account for EVO and remaining cylinder pressure and that changes with VE and rpm. That with added deflection induces un calculated valve velocity.

3V Performance wrote: ↑Tue Oct 29, 2019 8:04 am The problem is how do you account for EVO and remaining cylinder pressure and that changes with VE and rpm. That with added deflection induces un calculated valve velocity.

I’m not worried about EVO event in terms of exhaust valve bounce. The pressure is higher in the cylinder than in the port, the valve has been closed for a while, and the camshaft is trying to open the valve. I think that the valve and lifter are going to stay on the cam lobe reasonably well at that point. This is a DOHC with hydraulic direct acting bucket lifters, so I am thinking it’s pretty rigid and it’s not going to spring load the valvetrain the same way as a pushrod system. Is that a safe assumption?

I’m specifically worried about the EVC and valve bounce related to that event.

ptuomov wrote: ↑Tue Oct 29, 2019 8:25 am

3V Performance wrote: ↑Tue Oct 29, 2019 8:04 am The problem is how do you account for EVO and remaining cylinder pressure and that changes with VE and rpm. That with added deflection induces un calculated valve velocity.

I’m not worried about EVO event in terms of exhaust valve bounce. The pressure is higher in the cylinder than in the port, the valve has been closed for a while, and the camshaft is trying to open the valve. I think that the valve and lifter are going to stay on the cam lobe reasonably well at that point. This is a DOHC with hydraulic direct acting bucket lifters, so I am thinking it’s pretty rigid and it’s not going to spring load the valvetrain the same way as a pushrod system. Is that a safe assumption?

I’m specifically worried about the EVC and valve bounce related to that event.

I don't have any experience with DOHC systems on the spintron so I can't say for sure about system deflection BUT I'm sure some exist's. From my experience bounce is started from opening cycle most of the time on both intake and exhaust valves. Excessive loft ( late over shooting closing ramp )is also a player. Spring surge contributing to both scenarios most time from valve opening event.

3V Performance wrote: ↑Tue Oct 29, 2019 9:56 am I don't have any experience with DOHC systems on the spintron so I can't say for sure about system deflection BUT I'm sure some exist's. From my experience bounce is started from opening cycle most of the time on both intake and exhaust valves. Excessive loft ( late over shooting closing ramp )is also a player. Spring surge contributing to both scenarios most time from valve opening event.

Out of curiosity, how do you control that bounce that starts from heavily loaded pushrods etc. at EVO? Is higher valve open spring load and the spring installed the right distance from coil bind at valve open the only way, other than more rigid components such as pushrods? I guess slower opening exhaust valve might help some.

In any case, I believe that the DOHC direct acting bucket lifter should inherently be much less susceptible to this problem. Our main considerations are the V8 firing order and exhaust back pressure around EVC, in addition to the usual rpm related issues.

To simplify the question further: ignoring the exhaust gas pressures and generally the gas pressure differential between exhaust port and cylinder, are the required valve seated loads proportional to the reciprocating component weight and proportional to the square of the engine speed?

ptuomov wrote: ↑Tue Oct 29, 2019 11:29 am To simplify the question further: ignoring the exhaust gas pressures and generally the gas pressure differential between exhaust port and cylinder, are the required valve seated loads proportional to the reciprocating component weight and proportional to the square of the engine speed?

No
It is not as simple as that, everything matters and everything changes with each application.
dennis h

dennis h wrote: ↑Wed Oct 30, 2019 7:42 am

ptuomov wrote: ↑Tue Oct 29, 2019 11:29 am To simplify the question further: ignoring the exhaust gas pressures and generally the gas pressure differential between exhaust port and cylinder, are the required valve seated loads proportional to the reciprocating component weight and proportional to the square of the engine speed?

No
It is not as simple as that, everything matters and everything changes with each application.
dennis h

Why isn't it as simple as that? If you start from a known good combination with a known critical seated load, why wouldn't a small change in component weights, holding everything else constant change the required load proportionally? Why wouldn't a small change in rpm^2 result in proportional change in the required load? Where does the high school physics rule of thumb go wrong?

ptuomov wrote: ↑Wed Oct 30, 2019 9:23 am

dennis h wrote: ↑Wed Oct 30, 2019 7:42 am

ptuomov wrote: ↑Tue Oct 29, 2019 11:29 am To simplify the question further: ignoring the exhaust gas pressures and generally the gas pressure differential between exhaust port and cylinder, are the required valve seated loads proportional to the reciprocating component weight and proportional to the square of the engine speed?

No
It is not as simple as that, everything matters and everything changes with each application.
dennis h

Why isn't it as simple as that? If you start from a known good combination with a known critical seated load, why wouldn't a small change in component weights, holding everything else constant change the required load proportionally? Why wouldn't a small change in rpm^2 result in proportional change in the required load? Where does the high school physics rule of thumb go wrong?

You are using "weights" (plural) which implies changing more than one specific type or class of component in the combination. So really you are not changing one thing and holding everything else the same. Even holding the mass constant in a single component but changing the mass distribution can affect the resonant frequency.

Kevin Johnson wrote: ↑Wed Oct 30, 2019 10:31 amYou are using "weights" (plural) which implies changing more than one specific type or class of component in the combination. So really you are not changing one thing and holding everything else the same. Even holding the mass constant in a single component but changing the mass distribution can affect the resonant frequency.

I'm defining the reciprocating mass as a single mass in this thought experiment and including half the spring weight with it. Do you have any reason to believe that this is a quantitatively poor approximation for this problem in a direct acting bucket DOHC valvetrain?

IVO to mid-lift valve motion is controlled by the camshaft lobe. Mid-lift to max to mid-lift is controlled by spring force.

Spring force is a function of spring rate, installed height and valve lift. Since valve lift changes with crank angle, so do spring forces;
they peak at max valve lift.

Valve train force is a function of ensemble mass (valve, retainer, keeper, spring, push rod, rocker, tappet, etc),
and valve acceleration. Valve acceleration is determined by cam lobe shape and rpm.

Now things get interesting. Both spring force and valve forces will change with crank angle; valve forces are also
affected by engine rpm.

What is necessary is to plot both force curves against crank angle (at various engine rpm), to find the engine rpm at
which valve acceleration force exceed the spring control force. At the cross-over point, the valve ensemble will separate
from the cam lobe and the valve will float over the nose of the cam. When the valve floats over the nose, it crashes
down on the seat at closing, causing it to bounce. And bounce it does, often quite high off the seat.

The valve separation point, where spring control is lost, does not occur at the nose or at peak lift. It occurs at peak
valve acceleration crank angle. Knowing the actual cam lobe acceleration value is essential to determining required spring
force for any valve ensemble mass or cam lobe design.