of course pressures will equalize in a closed system and as they do, fluid flow will decrease
but while in motion a steam engine is not a closed system. steam is constantly exiting thru the exhaust ports of the cylinder and out the stack.
Part of the equalization involves the steam circuit design. In the years before the ASME ‘restored’ the locomotive-boiler code, there was some discussion whether a superheater needed its own separate safety-valve arrangement under Part 230 as a ‘separately fired pressure vessel’ – the logic being that with good long-lap valves at mid at one end, and a tight throttle at the other, carryover ‘trapped’ in the elements might build up dangerous pressure. Now this is true of a dome throttle, but a modern front-end multiple is ahead of the superheater elements, so any pressure increase that might occur in them relieves back along the dry pipe to the vast reservoir of overcritical water ‘under the dome’ so no explosion is possible.
Carryover into the tract between the multiple throttle and the valve chests is certainly possible, but that’s much more a bursting hazard in the liquid phase than it could ever be as a ‘fired’ vessel in the smokebox…
Something else that is fun is to see how much of the effect of very high boiler pressure is wasted with poor front-end design: it’s fine to get it into the cylinder faster, but then you have to get it out again after it’s volumetrically expanded…
Not so. When lifting heavy tonnage, the reverser will be ‘in the corner’, meaning the valve is slower in action and leaves the inlet port open longer so that full steam can work longer against the piston face. When a steamer is skipping along at limited speed, that is when the valve covers the inlet longer and restricts admission so that the steam has to work longer in the piston’s stroke.
i know a steam engine is far more complicated than my simple explanation below which explains the basics
when the throttle is first “cracked”, with the locomotive standing still, steam flows thru the throttle raising pressure in one or both cylinders (depends on valve timing). Flow thru the throttle decreases as pressure rises. At some point, there is sufficient pressure to move the cylinder piston.
when the piston moves, cylinder volume increases which impacts the increase in pressure. In an ideal world the product of pressure and volume is constant, PV = C, Pressure decreases as volume increases. This is only true for a closed system (no change in steam mass).
Pressure remains constant if the flow thru the throttle is equal to the rate of change in volume (constant density). Pressure decreases if flow is less and increases if greater than the rate of change in volume.
And any change in pressure will affect the flow thru the throttle.
If the flow is more than enough to accommodate the increase in volume, pressure will increase, applying more force on the piston and increasing the change in volume.
If the throttle is open too wide, pressure will increase rapidly and excessive piston force can result in slip.
as speed increases, there is more consistent steam consumption. The need for greater steam flow necessitates the need for both less restriction between the boiler and steam chest (more open throttle) as well as a greater pressure difference between the boiler and steam chest. (I believe this is analogous to Ohm’s Law where pressure is like voltage and steam flow like current).
if there is a need to increase speed, the throttle may be opened as speed and consumption increases. Depending on cutoff, cylinder ports will not be opened all the time and steam chest pressure will be allowed to recover (increase more so) when valves are not opened.
at hi
Disregarding leakage the system is closed until the locomotive begins to move. Theoretically, the steam only “leaks” out of the cylinder after the steam chest is again a closed system. How much can the pressure in the steam chest actually drop during the time the valve is open? Or does the overlap between the two cylinders (among the three?) effectively mean there is constant flow through the throttle into the steam chest and up the exhaust. Just posing the question answers it, never a constant flow, always a series of discrete lumps of steam. Hence the chuffs.
This thread seemed to me to be about what happens to pressure in the steam chest as the throttle is opened as contrasted to when the cut off is changed.
Pressure transfers virtually instantly, speed of sound in the fluid actually, which is pretty damn fast. The bangs you hear from unmuffled exhaust (or badly muffled as in a Harley V twin) are sound waves traveling down the exhaust pipe much, much faster than the gas flows. Pressure moves through a fluid much faster than volume can move (or bats and dolphins would need very hard foreheads).
For ICE piston engines using valves it is thought there is an rpm limit dictated by the speed air can move through the inlet and out the exhaust. At say 22,000 rpm, the maximum achieved by F1 V10 before the big rule changes to V6 turbo hybrids, engineers thought they might be getting close. Even then the gas flow is “particle” more than “wave” but wave effects have enormous effects. In theory a steam engine does not the need discrete volumes required by a piston engine, maybe that’s why the steam turbine preceded the gas turbine.
What peak rpm is reached by a high speed steam locomotive bearing in mind only two cylinders and related valve mechanisms are involved in the majority of engines.
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Your question implies both. The relationship between the throttle and the cut off should be the same whether you change the throttle but not the cut off or vice versa. The two devices affect the flow of steam into the cylinders. The throttle limits only the maximum possible flow. The cut off limits the more or less instantaneous flow of steam into the two cylinders.
Pressure in the steam chest should always be the same as upstream of the throttle. This should be true until the limit of the ability of the boiler to deliver steam is reached. In a well designed locomotive I can’t see this ever happening.
Now, why would that not be so?
I think some of the discussion will be more straightforward if you recognize what the throttle does, and where it is.
An American multiple throttle is located at the superheater header, and consists of a series of relatively flow-streamlined poppet valves that open sequentially via cam as the lever is moved. Since even this would be hard to move initially, the first poppet is purposely made very small, so there is less need to ‘hog’ the throttle open and then slam it most of the way shut for making smooth starts – I trust you recognize what about a dome throttle makes that practice less critical in producing excess torque. The older Wagner throttle (crca 1912) was an early and very effective form of fluidic amplifier, using a relatively small coaxial control valve to regulate differential steam pressure on the main throttle valve, and inspection of that patent and its claims will go a long way toward explaining to you what happens as steam traverses a partially-open throttle valve into a ‘draining’ load.
Keep in mind that as you open the throttle, the steam enters a fairly long, fairly cold tract, probably with high speed and associated flow and reflection shocks, and will suffer wall condensation to a higher degree until the tract comes up to reasonable temperature. (I am uncertain the degree to which hot smokebox gas actually heats this tract; it may be considerable, but one notes that all the economizers like the Pielock that attempted to use smokebox-gas heat for steam superheating were more or less functional failures). So you don’t have immediate boiler pressure ‘through’ the throttling orifice, and then again you don’t have immediate boiler pressure evident at the valve chest. Further condensation then occurs
The chest pressure only ‘drops’ because flow into the cylinder exceeds flow into the tract. At starting ‘cyclic’ it may actually represent a flow impediment as the volume has to be filled from the boiler before higher pressure will be ‘seen’ at the piston head.
The concerns with steam-chest pressure are, I think, much more relevant in the range of speeds that the time the long-lap valve is closed to steam cause pressure effects in the steam tract. I think of this as analogous to some of the effects with proportional relieving and recovery of compression.
But you have to answer that question as fairly as you asked it, and the answer is ‘no’ but only incidentally related to ‘chuffs’. Remember that you don’t hear the admission (this is why the South African 25 class is virtually silent with the fans turned off, and a Mallet working compound has only one audible exhaust signature) and the ‘chuff’, being a release to nominal back pressure, is not a comparable rate to admission mass flow.
One way to look at this is to examine a torque diagram like those in Wardale’s Red Devil book to see where the little transition at admission cutoff comes, then compare this to indicated pressure (which also reflects tract and passage pressure to within a small amount). But it is complicated becau
MEF is the “halfway” pressure on the piston face. Reducing the cutoff time will lower the MEF because the steam pressure is now stretched for a proportionally longer portion of the piston stroke.
Roughly speaking, cutting the cut off time in half should cut the MEF roughly in half. Throttle position should have no effect.
I looked a little at the superheater effect and much to my surprise the steam is superheated in expansion tubing so as the keep the pressure of the dry steam at boiler pressure. Adding heat can cause fluid flow without necessarily increasing pressure. At least, not by much. Of course there must be a pressure difference for fluids to flow but a steam engine boiler system is a closed system until the exhaust valve opens.
Steam expands from the boiler through the throttle valves into the superheater, expands into the steam chest and expands into the cylinder when the cut off controlled valve opens where it is trapped inside a chamber of changing volume. Only there should there be any appreciable pressure drop.
In steam it’s MEP (because measured by indicator rather than strain gage; it’s highly instructive to make the actual piston-thrust measurement, but this argument doesn’t contain that…)
And it is NOT a ‘halfwit’ pressure; it’s the integral of the instantaneous pressure over the length of the stroke from admission to exhaust cutoff. That’s a highly different thing because of the ‘double salience’ imposed by the rod geometry. If you wondered how an engine with an indicator diagram like, as Angus Sinclair so wittily noted, a small leg of mutton could show better over-the-road performance than one with a textbook-clean diagram – here is part of your answer.
In brief, even just with reference to a good indicator diagram (which if you’re lucky will actually have an indication of MEP calculation on it) … look at the rectangular hyperbola from cutoff to exhaust release? the MEP of concern is the line that produces equal AREAS under that curve to either side of it, and on a good diagram you can almost read that off or extrapolate it to get a working number. The nice part is that this adds to the portion of MEP you read off the admission, which ought to be nearly linear (see below) to calculate your overall effective pressure…
So far, you are correct, but unfortunately your observation is self-evidently correct. That you haven’t looked at it well enough yet is clear when you then say
Don’t be ridiculous. You haven’t even remembered that the period of admission up to cutoff is not only finite, but rather long; MEP during this entire part o
ideally when steam expands, PV=C, illustrated below for 10, 40 and 80% cutoff. i calculate the average pressure (MEP) across the cylinder cylce as 33, 77 and 98%.
at 10%, for example, steam consumption is reduced by a factor of 10, but MEP is only reduced by a factor of 3.
A “mean” is halfway by definition.
If you could take the necessary pressure readings half would be higher than MEF and half would be lower.
MEF isn’t calculated that way anyway. It’s a deduced number from data that are not in themselves pressure numbers.
As far as I know locomotive engineers have a boiler pressure gauge. There’s no pressure reading taken from anywhere else in the steam system.
An analogy to a garden hose was used earlier in this thread. I suggest the boiler system is more like a household water system. Municipal line pressure is limited by a pressure reducer at the main service tap. This would be the throttle. The valves would be the taps.
If the system is designed correctly there should be in practice an “infinite” supply of steam until the water or fuel runs out.
A Newcomen engine was an atmospheric vacuum engine. The piston was sucked down by cooling the cylinder. Watt’s not so brilliant addition was an external condenser. Both types didn’t rely on steam pressure to do any work. Trevithick was da man.
Superheaters run at boiler pressure. There’s no device to separately restrain the dry steam. I had understood that there were some locomotives with superheaters that placed the throttle between the boiler and the superheater.
isn’t that the median?
Ummm…no. A mean is a figure derived by adding the values in a data set, and then dividing that total by the number of data points, by count.
You’re thinking of the ‘median’, as Greg has pointed out. That is the value which has an equal number of data points on either side of it.
Yes, but you picked the wrong metric even if you didn’t mistake the technical term correctly. What you meant was halfway down the cylinder, and that’s just plain wrong. “I’m a lawyer, Jim, not a statistician!” … but for someone who values silently-sniggering semantics as much as you do so often, terms sometimes do matter.
Now you’re mixing your units. The issue is that “MEP” is that pressure that, applied consistently from admission to release (technically a bit different but it will work for this purpose) would produce the same OVERALL work as the varying pressure profile ‘in reality’. That says nothing about where in the stroke the ‘average’ pressure would be reached, although when you look at the actual events it’s not that hard, and as noted a good indicator will give it to you even at high cyclic.
Meanwhile MEF measured someplace logical, for example at the key between the piston rod and the crosshead, is likewise an average over the stroke; we won’t yet have converted it into wheelrim torque (which is really, I think, the ‘figure of merit’ that ought to matter here if the translation to ‘force’ is going to be made).
Which is why it doesn’t matter to indicated measurements on a steam engine (where it is nugatory, 'cuz the pressure measurements actually mirror cylinder performance, which ‘pressure’ data in a gas engine do much less well).
Yes, that’s wh
Yes, I did mean a mean. Not the median. As in the middle of the range of pressures. Or halfway down the range of pressures. Used instead of the actual range of pressures. As if the mean pressure was exerted for the entire stroke of the piston. A derived number rather than an actual mean.
It’s a bit weird because the MEP isn’t actually calculated as a mean. It’s derived.
Abd of course MEF was a typo.
Yes. The mean in this case is not actually calculated anyway.
And I meant halfway down the pressure range, roughly.
I’m waiting for an intelligible articulation of the answer to the question posed. If we’re lucky we might get the how and why as well.
And one cannot silently snigger, however much one might be tempted. Nice play on “halfway” though, assuming that was intentional. But even a blind squirrel finds a few nuts.
Semantics ain’t easy.
looks to me that it’s the average pressure (mean) during the cycle. see Method of Ordinates which is what i calculated in my plot as well as using calculus
1448
Since perhaps the whining English schoolboy claims to be resolved on creeping like snail reluctantly to knowledge, let’s revisit those original questions:
As discussed, this is complex, probably has an answer that is time-variant (and cyclic-speed variant) unless one develops an analogue of MEP for steam-chest pressure, and already contains one semantic whopper in steam being ‘drawn’ into the cylinder – although this is really ‘critical’ only in understanding how the physics of flow will work. [The only time steam can be said to be ‘drawn’ into a cylinder is in the context of certain types of drafting/snifting, where ‘displacement steam’ is admitted just at atmospheric pressure to keep oxygen out of the steam tract and its vulnerable tribology, and even there a drifting throttle customarily provides a slightly higher pressure to ensure no air ‘leakage’…]
Steam is always ‘pushed’ into the cylinder (more saltily 'mass flow is achieved through pressure differential) and it should be clear that EP, as an index of extra table work via piston thrust, will be lower than chest pressure if mass flow during admission is to take place. Now EP at the piston face is partially determined by flow through the passages, so it may not necessarily be ‘higher’ than chest pressure, but any (necessary!) tendency to flow reversal at any such time would be a Bad Thing for subsequent expansive work. So in all probability a situation in which chest pressure cycles higher than passage EP would be a poor thing to be in initially.
If the cutoff is shortened (we have amply seen by now the point I keep trying to make about using clearer terms for cutoff) with the throttle wide, the chest pressure will now peak a bit higher, closer to boiler