throttle vs reverser/cutoff?

[gregc]

how close to boiler pressure is steam chest pressure or MEP cylinder pressure at max speed on level ground?

can this be quantified if we know the rolling resistance at that speed (see Armstrong's charts)?

Overmod

gregc is wondering if the transient rise in steam-chest pressure

no, steady state

1496

[Overmod]

Lastspikemike

Does the superheater act as a kind of buffer rising and falling in temperature in opposite phase to any pressure effects?

Not to my knowledge.  Remember that it is open at the 'back end' via the dry pipe to the ebullient zone (usually in the convection section) under the dome, so any tendency for pressure rise in the elements is largely balanced (it turns into lower density instead, another Tuplin plaint).

To the extent that the steam loses density as it gains temperature, it might be considered a 'buffer' but I don't think there are analogic benefits from considering it as such.  

I'm running around this morning but I'll give a little more thought to the idea when I can.

[Overmod]

Lastspikemike

It occurred to me that the pressure change in the steam chest due to opening the valve to the cylinder would be proportional to the volume added by opening the cylinder to the steam chest.  It looks to me like that would be quite a small effect.

It is small in terms of a number like lb/hr overall mass flow, but it is important as you get to higher speed and shorter cutoff where the physical flow into the cylinder over the course of admission becomes more and more physics and time affected.  One thing that became noted was that maintaining a reservoir of relatively high mass adjacent to a large (and quick-opening, cf. circumferential port opening in long-travel valves, the use of Trick or Willoteaux valves, the better sort of quick-lift poppets) and ideally low-volume but well-flow-streamlined passage that directs the flow as well across the face of the piston as possible, was a functional advantage in a critical part of early admission.  This is a thing that proper Carter-style reversible compression control simplifies (by having high mass at nominal pressure equalized across the piston face at the start of admission).

The ratio of exposed cylinder volume to total steam volume back into the boiler (throttle/ regulator unchanged in opening amount) surely must be small.

It is not so much that it is small as that it can be functionally flow-restricted, a lesson that repeatedly bit the Franklin people in practice.  Very little flow resistance is necessary to hold up the small early mass flow for good cylinder filling in the rest of admission (then, plus expansion in cutoff)

More so if the cutoff restricts that effective volume.

The principal evil is not so much 'volumetric' I think as the smaller effective port opening as the cutoff is shortened -- this is another thing that the long-travel valves can help with -- and the other impediments to 'mass flow' that trade differential pressure for velocity (which is of little good if it results in lower density in the charge steam).

Note that the turbine has a clear advartage here in addition to continuous flow: any pressure-to-velocity can be captured in the initial impulse stage, then expansion through the subsequent reaction stages uses the expansion energy.

Once the cut off closes the steam entry the volume increases by movement of the piston, heat is converted to kinetic energy.

But what's of more interest here is the things that happen on the other side of the cutoff valve, since that is now mechanically divorced from the cylinder expansion.

In a good long-lap long-travel setup you'll get throttling of the admission flow as the steam edge approaches cutoff, but with a good valve this can again be quick -- something to note here is what happens to the flow from the boiler through the throttle restriction gregc posits through the tract and steam chest to the port area when the valve shrouds and then closed.  I had to argue with people discussing 'momentum' effects in the steam but there is clearly a mass in motion that has to be 'decelerated' -- it just does it with gas kinetics and not elastic/inelastic impact.  I will not be surprised to see reflected shock in various places, perhaps even flow reversal, and this may become more than usually important at high cyclic if any of that 'action' persists at the next admission.  Now, in this context... 

the kettle upstream of the valve maintains the pressure in the steam chest at the previous value.

Except that a better way to put it might be tries to maintain the pressure -- gregc has intentionally made the throttle restriction great enough to retard flow in admission, or there wouldn't be a question, but low enough that a net amount of pressure recovery, and corresponding density close to the valve, will be available at the moment of admission, but not supporting the mass flow which that pressure would produce if it were actually 'kettle pressure' at full kettle generating mass flow.

Put another way, since the volume of the steam from boiler to valve remains the same throughout...

... but it doesn't, in this constrained situation...

...to gain pressure would require a temperature rise throughout the volume.

I do not think this follows: the pressure rise past the presumed throttling restriction is a flow equilibration, and if the tract is heated and reasonably well insulated, as it would likely be in sustained running,  the superheated steam should heat very little as it is pushed into the now-closed volume of the steam chest.

The superheater does not increase system pressure so it is hard to see how closing the cylinder valve could.

Here, because it is no longer freely linked to the boiler; it is on the other side of a throttled opening.

Opening the throttle more would not change the pressure.

But by reducing the restriction, it allows 'quicker' as well as more pressure rise at the chest volume both during admission and in rise to equilibrium pressure after admission has ended.

Closing the throttle some should not either.

That's common sense, because 'even more' throttling restriction over what gregc had already will further cause starvation pressure drop over admission and (probably) limited rise in 'pressure' (here, mass ready to flow at next admission exerting that higher pressure).  This is the area where the question may give a counterintuitive answer to observed practice...

The selected throttle opening would likely be fully open  most of the time.

In part because that 'removes all doubt' that the best possible mass flow down the tract will be occurring at any point over a cycle.

[Overmod]

gregc

no, steady state

But it has to fluctuate, by definition, for your effect to be real at all, and it's a reasonable example of why you might want to indicate the chest and passage pressures as well as cylinder pressure under some conditions.

If the steam chest pressure 'recovers' during the cutoff time, either it produces the same 'loss' in admission as you'd see from the tract, or there is some restriction in steam flow sufficient to starve admission MEP.  The second HAS to be the case in your example, and moreover the 'reservoir' effect in the chest volume needs to be large enough that what enters during cutoff exceeds the mass flow from the chest down the tract during admission.  Outside that envelope there is obviously falling horsepower on the low hand, and lower pressure loss over admission on the other.

[gregc]

of course it fluctuates, it drops during admission and recovers in between.   it drops less when cutoff is reduced.

but the average pressure is higher

[Overmod]

gregc

but the average pressure is higher

I keep saying this. IT DOES NOT MATTER what the 'average' pressure in the chest is.  It could recover to within range of boiler pressure within 15-20 percent after cutoff... WHO CARES??? The steam sits there pushing on nothing meaningful until next admission.  Then its "pressure" starts high but likely starts falling away rapidly as the passages fill and the admission phase begins.  That rate will be based on expansion of the various zones back to your throttle 'orifice restriction' (it is shaped differently but functions the same in this context.

And falling admission leads to lower EP at the moment of cutoff, after which no one cares again what the steam chest does for a while.

[Overmod]

Lastspikemike

And if it cannot recover?

Look at the opposite of underdamped resonance in control theory.  What will happen is that the peak chest pressure will be achieved later and later in the expansion 'dead time' until it actually starts to stay 'drooped' by the moment of admission.  Less mass will produce unquestionably lower power after that, and what is likely is that the engine will reach an equilibrium at lower MEP that of course implies lower speed ceteris paribus.

If each time the cylinder is opened to the steam chest the pressure in the steam chest or indeed the whole system drops can it keep dropping?

This is a pipeline-has-two-ends problem, with perhaps an easier answer than you realize, because the cylinder volumetrics are speed-dependent but the throttle opening isn't.  You have the throttle partly shut so only a given mass flow gets through, and the pressure differential across that effective flow restriction and the free area across the throttle opening will determine what that is.   Assume for a moment that the steam-chest volume adjacent to the inlet valve has 'bled' up to effective equilibrium with the tract at the moment of admission.  Steam rushes into the fixed volume of the passages, then begins to move the piston and the cylinder 'begins to take steam' proportional to piston movement. If there is 'not enough flow' to keep the admission up, the pressure may start high but fall off quickly as admission proceeds.  At the 'point of interest' the MEPs from moment of admission to moment of cutoff could be thought of as equal ... but with different indicated diagrams ... for the case where the throttle is closed 'too much' but pressure recovers intermittently vs. where there is just enough coming down the tract in average flow to produce the same mass flow by the moment of cutoff.  (In practice this would be complicated by rod angle, etc. as the torque is not linearly related to piston thrust -- there is a good graph in the Red Devil book that shows resultant torque at 15-degree intervals for a good unrestricted 2-cylinder DA engine, and you could probably get an adjusting factor from that to get a reasonable ballpark approximation of performance difference.

The superheater still has me puzzled. The hotter the superheated steam gets the more volume it occupies. Pressure doesn't rise. If a low pressure wave travelled from the valve opening back to the superheater tubes then the superheated steam would cool a by a small amount. It should be heated almost immediately expanding as it does so cancelling out the pressure drop caused by the valve opening to the cylinder.

But the outlet to the superheater is very far from the valves, so there may be low resonance effect, and the throttle is interposed giving a flow restriction to shock or very fast flow changes.  In practice I do not recall seeing standing waves or other behavior in element flow... but I have not looked carefully for it, either.

Conversely, when that valve shuts, particularly when due to a shorter cut off, the resulting  wave should try to compress the superheated steam resulting in less instantaneous heat absorption. The buffer effect imagined might actually occur. Net result would be a tendency to cancel the pressure oscillations  that are being considered as possible.

Without running numbers, I think that any oscillations surviving reflection at the return bends would be damped relative to the overall mass flow through the element.  You might look at the research to see if there are resonant effects there in English superheater arrangements that open more directly to good flow-streamlined tracts.  I am still thinking about the physical effect of the less dense but faster-moving flow at the superheater on any reflection due to starting and stopping flow quickly down at the valve.

[gregc]

just to be clear, the cylinder valves sit in the steam chest or are fed directly from it thru a large opening.

Overmod

WHO CARES???

isn't it the starting pressure of the steam entering the cylinder?

[PM Railfan]

"just to be clear, the cylinder valves sit in the steam chest ...." - No. There are no cyclinder valves. They are called "valve cylinders". One each side.

Yes, the 'valve cylinder' resides in the 'steam chest', right above the 'piston cylinder'. The 'valve cylinder' is named so, because it is a cylinder, that is used as a [sliding] valve. Not cylinder valves, that is a whole other type of valve.

Your picture is wrongly termed. "Steam chest extends across twin valves" makes no sense. The "Steam chest" is not a pipe. It is the chunk of cast metal that houses the  valve cylinder AND piston cylinder AND contains the porting neccessary to operate the cylinders.

As a clothes 'chest' holds drawers, a steam 'chest' holds cylinders. It also holds the "yoke" or "saddle". Thats the part the smokebox rests on, on top of the front of the frame.  

 

"Does the superheater act as a kind of buffer rising and falling in temperature in opposite phase to any pressure effects?" - NO, the temperature of the superheater is controlled by the firebox. And is used to 'super'heat the steam. Its a controller, not a controllee.

You dont want the superheater temp to be able to fall. If anything the whole idea of a superheater was to rise, and stay as hot as you can get it!

The formula goes: higher the heat = hotter the steam = more work can be done.

 

PMR

[Overmod]

PM Railfan

Your picture is wrongly termed. "Steam chest extends across twin valves" makes no sense.

Fascinatingly, it makes perfect sense if you recognize what that picture shows. Remember I mentioned Wardale's design that used separate piston valves (of his very interesting detail design) for inlet and exhaust?  This is the same arrangement of valves, side by side across the top of the cylinder a bit like the paired poppet arrangement in some Franklin variants -- I think here intended only to give more effective port area or whatever with the two valves driven in parallel, but I don't see either the valves or the valve-gear arrangement.  That the "steam chest" is common to the valves indicates that inside admission at least is in parallel.  Note what appears to be colossal outside tract arrangement at each end, perhaps to help exhaust clearing.

Perhaps now would be a good time to note that this is HIGHLY atypical setup, not at all characteristic of a normal piston-valve arrangement.  (It took me a moment to recognize I wasn't looking at a Wagner bypass-valve cutaway (see for example the restoration pictures of ATSF 2926) as the 'twin' arrangement is so unusual...)

gregc, can you provide a link to the other 3D detail pictures of that arrangement?  I'd like to see whether those will be articulated valves with many thin rings, and how the four of them will be driven.

The "Steam chest" is not a pipe. It is the chunk of cast metal that houses the  valve cylinder AND piston cylinder AND contains the porting necessary to operate the cylinders.

well, no.  By definition anything after the steam edge of the valve in port and passage is... well, ports and passages, dead space, nothing to do with a 'steam chest' at all.

"Does the superheater act as a kind of buffer rising and falling in temperature in opposite phase to any pressure effects?" - NO, the temperature of the superheater is controlled by the firebox. And is used to 'super'heat the steam.

In all fairness, he was discussing the flow of steam in the inside of the elements, not the gas flow over the outside or the heat-transfer characteristics across the element  thickness to the relatively good insulator that superheating steam represents.

I am not sure where that 'controlled by the firebox' business came from.  Superheat is a consequence of combustion-gas flow.  If you are not aware of superheater dampers ... and why later designs stopped using them ... I recommend that to your attention.

You dont want the superheater temp to be able to fall. If anything the whole idea of a superheater was to rise, and stay as hot as you can get it!

There are VERY good reasons to keep the degree of superheat limited even in terms of what relatively conservative element arrangements produce in normal operation.  In practice you would likely want to design the degree of actual superheat in a reciprocating locomotive to be little if any more than that which precludes wall and nucleate condensation -- those being the purpose of having superheat in the steam rather than higher saturation pressure.  (And while you can tolerate more in turbines, there is not a proportional gain in expansive work purely from the added superheat above that which precludes nucleate condensation in the last-stage blading that offsets the higher overall cost to generate it.)*

The formula goes: higher the heat = hotter the steam = more work can be done.

Up to the point your valve (or throttle!) lubricants carbonize, or your spool sticks in the cage and seizes at 9rps, or your elements start to soften or tips start to burn.  In practice you can only get much above 900 degrees in an oilless engine like a Spilling, or with some care in an asynchronous compound (which has no contact piston rings) and even then you have to start doing steam cooling a la Wardale on things like the glands as well as the, well, ahem, parts of the cylinder block that Wardale steam-cooled to preclude effects of excessive superheat.

I suggest you read up on some of the characteristics of highly superheated steam -- and the practice of desuperheating in power-station operations -- before pontificating that moar superheat is a design principle.

*If you can find any of the discussions of the original enginion AG ultrasupercritical engine design, you can read about a version of extremely high superheat (950 degrees C being the number that first caught my attention) produced in a separately-fired superheater for use in a 'positive-displacement expander' (really just fancy jargon for a piston engine).  This design could use high superheat because it metered (as in modern low-fuel-mass direct fuel injection) a very small amount of (again going by memory) circa 7550psi USC steam with pilot-injection techniques, supposedly long-expanding it down to where it could be expediently recompressed after long expansion to go back into the BFP feed train without ginormous heat-exchange condensing.  See if the 'ZEE engine' technical paper is still up; that was a somewhat less extreme but still fascinating version of the technology, and (if you believe some of the hype) was slated to become a reliable tech for distributed electricity generation, 5kW per unit...

[gregc]

Overmod

gregc, can you provide a link to the other 3D detail pictures of that arrangement?

https://www.advanced-steam.org/5at/technical-terms/steam-loco-definitions/steam-chest/

... getting off track

[Overmod]

gregc

... getting off track

Far from off-track, as that page will serve usefully to inform PM Railfan on what the 'steam chest' does (much more a treasure chest than a hope chest!) and on what its purpose in an optimized design would be, according to the opinions of some very wise people.

Note that to Wardale it's almost a truism that the 'volume' throughout the steam chest not only have 'recovered' from any transient pressure drop during admission during the subsequent period of cutoff but that it represent a reservoir of high-pressure steam equal to (or as noted 'greater than') the swept volume to be filled during admission.  (The peculiar thing is that at least the earlier version of the 5AT collateral implied that in SGS that followed the FDCs the chest volume should equal the full swept volume, which is unnecessary with even reasonable tract-flow streamlining down to the chest volume.

Inherent in the explanation is the idea that there be enough volume aa near 'boiler pressure' as possible to get admission filling -- you will have recognized that this implies prompt and complete mass flow down the tract to 'make up' the volume as it transfers out of the 'reservoir' and that it do so with enough near-instantaneously-communicated pressure to keep the EP from falling at the piston face (that being net of energy extracted via thrust during the admission period, when any 'expansion' in the steam leaves the equivalent of money on the table at the instant of cutoff).

In that context you can see the relative artificiality of discussing a steam-chest pressure that "fluctuates" downward during parts of the stroke.  For it to actually drop during admission is an indication that either the throttle is artificially closed or that the tract is defectively restrictive.  I don't have my copy of Carpenter's translation of Chapelon's handy, but a great deal of the theory behind 'Andre the Giant's' theory of steam 'streamlining' was actually reducing impediments in mass flow to the valves that accomplishes the necessary pressure-maintaining through admission with no need of recourse to a local 'reservoir'.

If I recall correctly Porta explained a bit about the use of Wagner throttles (he charmingly spells them 'Waggoner' and does not explain what they are, classic LDP) in the patent description of the ACE3000.  The idea as stated involved provision of receiver volume very close to the valves (which in a full proper high-speed design would be adjusted via IP injection to full design pressure at least over the commanded range of admission) with little flow-smoothed proportional trim throttles as near the ports as desirable.  (A variant of this system is one of the correct answers to high-speed slipping on unconjugated or clutch-conjugated duplexes, but we can take that point up later as it is a digression here.)

The moral, to the extent there is one, is that 'fluctuations in steam-chest pressure during admission' are a bit like the old UNIX International joke that if you see a command-line interface you've failed utterly as a systems designer.  To the extent there is any droop in indicated pressure at any point during admission, some of the available power is being lost, and unless full pressure (or at least as high a pressure as possible) is achieved in the cylinder at the moment of cutoff, the 'best' efficiency against somewhat fixed 'exhaust resistance' will be compromised from 'what it could have been'.

While it doesn't help the steam-chest thermodynamics, I'm going to reiterate why net decrease in MEP at moment of cutoff can be valuable: a number of effects both in propensity to slip/spin and to sustain slipping for longer degree of rotation per 'incident' can be lessened with restrictive steam-tract conditions alone.  In a slip the rate of volumetric increase in unit time increases, just as it would for the higher cyclic associated (with necessarily higher mass flow against greater 'piston resistance') at high speed, and therefore an 'automatic action' that causes admission MEP to fall at such a moment will assist in recovering the slip.  I would gently suggest that one of the lessons taught by a Franklin type A equipped duplex is that perfectly high chest pressures can really become annoying if you can't trim for chronic slip...

[gregc]

Overmod

Inherent in the explanation is the idea that there be enough volume aa near 'boiler pressure' as possible to get admission filling

gregc

how close to boiler pressure is steam chest pressure or MEP cylinder pressure at max speed on level ground?

can this be quantified if we know the rolling resistance at that speed (see Armstrong's charts)?

at speed, i assume average steam chest pressure is far from boiler pressure

1881

[Overmod]

Incidentally, here is the time to put in a plug for the 'full version' of the FDCs as produced by the 5AT organization and (AFAIK) still available from the successor-in-interest.  There is probably no better overall reference on practical design of a locomotive, and while it is likely still 'expensive' (Wardale has stated reasons for that) and may need some judicious interpretation for use in a North American context outside 5AT size, you should have -- and pass along when time -- a copy.

There is still not, to my knowledge, a version of Charlie Dockstader's 'valve gear on the computer' that extrapolates the geometry to CFD-derived steam-flow changes as it is adjusted.  That does not diminish the goodness (or continued usefulness!) of those programs, which were ported at least through early Windows to XP compatibility (and will still run effectively in emulation).

While you need certain acquired acumen to interpret the results, Bill Hall's programs (including Perform and Perwal) are highly recommended as tools.

 

[Overmod]

Here might be the point to take up the other half of the discussion: are there circumstances where interrupted flow produces an effect analogous to pulse tuning (previously alluded to here and a few places elsewhere) and, perhaps, be so tuned as to cause an effective pressure rise in the steam chest during admission that helps 'snap' effective mass flow as admission times become shorter and shorter and transient flow restrictions more and more significant, especially at very high cyclic?  That the effects are present in obtaining higher mass flow in IC reciprocating engines is beyond question.  I have not seriously looked at this (primarily because there are many ways to get it wrong vs. any that might be effective, and much of the required technological complexity is probably unsustainable in service at any reasonable price, let alone be cost-effective.  But do not let that interfere with the fun.

[gregc]

Lastspikemike

If the volume of each side of the steam chest is equal to or greater than the swept volume of the cylinder it feeds then that should be more than adequate to maintain pressure at say 50% cutoff.

~~bingo!    so how would steam chest pressure be affected if cutoff were reduced to half?   or if only half the volumne of a cylinder needed to be filled from the steam chest/

[Overmod]

Lastspikemike

If the volume of each side of the steam chest is equal to or greater than the swept volume of the cylinder it feeds then that should be more than adequate to maintain pressure at say 50% cutoff.

Sorry to keep harping on this, but it's only the swept volume of the part of the cylinder up to effective cutoff that 'matters' to steam-chest pressure fluctuation.  Once the valve is shut there is nowhere for the steam to go for a while...

Plus the two cylinders are 90 degrees out of phase. Pressure will be available from either of the two connected steam chests.

No, not particularly; the two steam chests 'in quadrature' are not even remotely connected to each other; even where there is only one throttle outlet, the branch will be high in the smokebox and the junction flow-streamlined for mass flow to the cylinders, so decreased flow down one branch will only serve, at most, as a sort of fluidic switch inducing (slightly) more flow in that short time to the other branch.  If you had one pipe going to what now would need to be called a Moor's head, and short branch pipes from that to the chests, you might see more of a rise, but I know of nothing built that way that was ever as successful as an engine with separated and flow-streamlined tracting.

From a pressure perspective each cylinder is supplied by both steam chests...

If there is any pressure effect at all, it will be reflected pressure or shock.  There will pretty obviously be negligible actual mass flow 'from one chest to the other' no matter how black you draw the arrows...

Due to the efficient design a two cylinder locomotive is really a four cylinder.

Well, no, it's a two-cylinder double-acting, which is no more a 'four cylinder' than an EMD 12-567 is a twenty-four cylinder.  Not only were there double-acting IC engines (the delightfully dotty Doxford providing some wonderfully Britannic examples) but there were attempts (one of which essentially bankrupted the fine old British firm of Kitson's) at using gas power on one side of a DA cylinder with steam of some quality on the other -- you can start rotting up all the disastrous why nots almost as soon as you hear the idea, but Still carried the torch to no few ultimately resentful and rueful folks...

I'm reminded of the 8 cylinder FM opposed piston diesel which has 16 pistons...

This is like the contrapositive of a DA... especially when you realize there is no objective reason an OP engine can't be DA as well (and yes, there were people who thought that was a honking good idea, and yes, Bill Gates did support the OPOC development financially... but I shan't digress further)

[quote]Point being that each steam chest fills two cylinders to the same cutoff level each revolution of the drivers.[quote]But we take this into account because the period of concern is over a stroke, not revolution of whatever the piston thrust is driving.

Part of why this is important is that there is part of the stroke that y'all are ignoring, and that I've alluded to but perhaps not described explicitly enough to recognize.  This is the whole part of the 'first' stroke -- let's arbitrarily say the one going to BDC --after the practical point of exhaust closure (which, remember, is affected both by limited and shortened cutoff with piston valves and most radial gear) back to BDC.  There is a residual steam mass remaining in the cylinder at this point, and it stays there to the extent the piston seals properly; it is critical to practical engine operation, in fact, that it is there.

Now take the 'second' stroke, still considering the engine as non-self-starting.  Admission to the cylinder ideally comes with the piston still very close to BDC, but the valve has been moving for some time already (the magic of long-travel) so the ports come open, the steam flows in for the admission profile (note in most engines this isn't symmetrical as the piston rod, etc. is present in the back volume but not the front unless tailrods are fitted) up to effective cutoff, and the expansion then occurs between cutoff and release isolated from the chest as we have seen.  What is still there, though, is that steam mass from the previous stroke -- it is still there, and it becomes more and more compressed as the piston moves forward.  This compresses the steam in the dead space at the forward end, too, and it is not hard to see that at some point the EP of compression will equilibrate to the EP if expansion on the other side.  From that point physical work has to be done on the piston to get it through FDC, and there is utterly no guarantee that the peak pressure right around FDC won't rise substantially over the nominal boiler pressure -- in fact it was common for such high compression and its effects to be observed, at least to the degree that Okadee made a product line to mitigate problems from it.  Now one of the principles of engine design is to use this compression to decelerate the reciprocating part of the mass; it's been argued roller-bearing rods wouldn't be practical without it and even balanced conjugation doesn't address inertial stress in the mains.

Now consider what happens at the port and the volume in the chest adjacent to it when the steam edge of the valve cracks open to start admission -- with many psi of differential pressure for a moment going the wrong way.  That this screws up establishment of smooth admission mass flow for at least a couple of milliseconds goes almost without saying; when the entire practical time for admission at high cyclic and short cutoff is already only measured in milliseconds you can realize the potential horror...

And here the point of Carter's reversible compression control becomes more clear.  If the compression pressure is at chest pressure at the moment of admission, the communication of pressure from chest steam to all parts of the piston will be immediate, and the subsequent mass flow (if there is to be no drop in chest pressure) will be exactly the same down the tract as is going through the port for the duration of admission.

Obviously you want more reserve at high pressure close to the valve if there is inadequate effective mass flow for makeup; equally obviously you want as little flow restriction down the tract as possible.  And this brings us briefly back to using sliding-pressure firing rather than throttle flow impairment at controlling admission EP/MEP and hence power per stroke...

The two steam chests feed four cylinders of capacity.

This is true, but you'd get more by saying each steam chest feeds two cylinders of capacity -- and is influenced by what happens in those cylinders only.

At high speed the cutoff will be shorter than when accelerating to speed.

But watch it!  There will be a point of minimum cutoff, probably associated with best water rate on testing.  But when accelerating to higher speed the cutoff required will begin to be longer and longer as resistance requires more and more cylinder power -- acceleration just requires even more to reach higher speed.

This is where the magic 'low 40s' number comes in -- there will be some point where no more steam mass can be admitted to make better power, and here acceleration falls to zero and the train will be running at maximum achievable speed.  But that will be far from shortest cutoff.

In practice the shortest effective cutoff with controlled timing and duration (but no pilot or divided admission or modulated injection) was about 15% for British Caprotti, which is both precise and reasonably fast-acting without spring-return issues.  It is notable that the valve arrangement with by far the lowest dead space and lowest port restriction of a reciprocating double-acting design (Bulleid's original sleeve valves as applied to Highland Point and the Leaders) showed no thermodynamic advantage (and in fact Bulleid reverted to piston valves on the follow-up 'turf burner'.  Some of the theoretical design work by Doble and the Beslers indicated that very short cutoffs (in the range touted as mechanically achievable with the evolved British Caprotti, about 3 to 5%) will only be of use for substantially higher pressures than are practical at large-locomotive scale or noncondensing configuration.

Lowest steam chest pressure would be at maximum valve opening duration.

I am not confident enough to say this.  Very likely it might be structural, related to some unshrouding concern or physical passage or chest characteristics.

If we make the assumption that steam mass flow through the tract is insufficient to sustain steady chest pressure at the ports (as in gregc's example with partial throttling) then progressive droop in admission might be observed, but the droop might not be linear.  Remember that piston speed is not linear with stroke progression, and during the entire range of practical cutoff the admission volume will be increasing over time... so any deficit in mass flow will produce lower admission pressure as expected right up to shrouding ahead of cutoff.

I assume this drop would not be significant given the volume behind the push of steam into the cylinders.

BUT but but this whole question revolves more or less directly about there NOT being enough mass flow, hence volume or pressure, to keep EP up to chest pressure up to sustainably stable percentage of boiler pressure or effective best throttle pressure.

One of the classic approaches that 'might' be used is a variant of desuperheat: inject a small mass of overcritical water (for argument's sake, distilled or demineralized water 'floating' at boiler saturation temp and pressure) into the chest where it will not enter the ports.  This would flash at 'lower' pressure and then take up some of the superheat (at a time in the stroke it can't be used) to keep EP high at the period through shrouding and cutoff.

The shape of the steam chest castings in the photo at the link provided (UK practice I note) seem patterned for convenience rather than from any knowledge of gas flow. 

It was my opinion that the work done on the 5AT was some of the most careful from a gasdynamic perspective; certainly if it is derived from Wardale's contemporary practice the tracting is probably as well flow-streamlined as it can be without compromising other aspects of a practical working locomotive (Wardale discusses this in the Red Devil book and, I believe, in some of the material associated with the FDCs).

Sometimes functional streamlining doesn't 'look' particularly elegant (Jaray bring one of the masters of that... and do you remember the 'banana car'?) but still highly efficient.  That would certainly be true of the shape of all parts of the exhaust tract through the nozzles and diffusers through the shape of the airfoil at the bottom of the petticoat (Wardale in fact having instructed me on reasonable flow shaping to direct combustion-gas flow from disparate directions around to be coaxial with the steam-jet flow for entrainment...)

[gregc]

Overmod

Sorry to keep harping on this, but it's only the swept volume of the part of the cylinder up to effective cutoff that 'matters' to steam-chest pressure fluctuation.

yes, and that's why i ask about the impact on average steam chest pressure when cutoff is reduced

[PM Railfan]

"I'm having a hard time accepting that ...." - Dont worry, I am too. About most all the incorrect info being given here. Ill wait till the dissertation is over, then fill in the blanks with the correct answers, again.

PMR

[Overmod]

Lastspikemike

My reference to the FM engine was by analogy. In the case of FM opposed 8 the engine is actually 16 cylinders. Indeed the makers thought so, hence 16 44.

This is hysterical.  The 16 is the horsepower, not the 'number of cylinders'.  The man who designed the engine says they are single cylinders, so you have zero valid claim regarding the 'makers'.   The liners constitute a single cylinder.  The whole point of the FM OP engine is two opposed pistons in one cylinder.  'Two cylinders' with trunk pistons would require two cylinder heads, too.  One of the important points of the FM engine is that it does not have either cylinder heads or valve gear; the lenticular squish is provided mutually in the piston crowns. 

Now if you really need to ease your confusion by thinking of the swept volumes of the trunk pistons as separate 'cylinders' that is fine; just don't go confusing the people who need to know the actual design principle and detail design compared to engines with discrete cylinders or power assemblies.  In some other OP engine designs there may in fact be more support for your analogy; you could think of the Deltic as three V engines arranged so all the pistons wind up in mutually-opposed sets, and then the analogy makes better sense if an engine of that configuration were in fact assembled with cast monobloc cylinder-and-crankcase units 'en V' like any good 60-degree V engine, assembled so there are physical joints between them where heads would go on a regular V engine.  On the other hand if the cylinders are in one block and bolt to the crankcases below the swept area for assembly (and I think that mechanical triumph the Commer Knocker was made that way?) you'd still be better off with the two-pistons-in-one-cylinder interpretation.

In my opinion you might as well use the analogy that a Stumpf uniflow engine has two cylinders, since the ends of the trunk piston sweep little common wall area.  I prefer to treat this as one long cylinder with a long DA piston in it, but if it's easier to understand as two I'm not going to insist.

The point about the pressure excursions in the tract is easily addressed by indicating several points with fast-acting transducers and then examining the resulting data relative to a common (fast-clocked) timebase.   I have not heard that the phase of intermittent flow in one branch of the tract materially affects filling during admission in the other, but that does not make the idea 'wrong' and indeed other assumed effects in steam distribution have proven historically unjustified.  It might be interesting to look at some of the more extreme Q2 tests, which had some tract and passage pressures measured continuously, as it might be possible to deduce results.  

[Overmod]

PM Railfan

About most all the incorrect info being given here. I'll wait till the dissertation is over, then fill in the blanks with the correct answers, again.

When you get correct answers it will be interesting to read them.

Do keep posting; you might have become confused by some of the strange hypothetical things in this thread, and been judged unfairly.

[Overmod]

Lastspikemike

Well, the common V8 would be one cylinder then since all eight holes are bored in one chunk of metal.

Obviously not, because eight separate holes, one per trunk piston.  This probably (and yes, that's a Cartesian weasel word) carries over to the situation where two blocks -- for the moment let's use straight-eights as the example, as two 'do the business' -- are opposed to form an OP engine.

Say for argument that we take a couple of 210As out of early Zephyr-style motor trains.  These already have the scavenge ports, belt passages, even proportional scavenge-blower capacity to work as desired for an OP compression-ignition engine; it will not change anything in the prospective 'nomenclature' to replace the two blowers with one driven preferentially off one of the cranks (as in Stonehocker's 38 ⅜ engine).  Now we need to replace the power-assembly cylinder heads, valves, and central direct injector with ... something.  The first significant detail emerges at this point: the engine manifestly retains the separate liners as well as structure of the 'cylinder' part of the 201A (i.e. 'power assemblies')  These are frankly obviously 'cylinders' in Winton parlance, and that would and perhaps should carry over to describing the engine thus far.

 The FM engine features (and it is repeatedly described historically as a feature) a one-piece liner forming the effective 'cylinder' that the pistons run in and seal against; we do not have that yet in our hypothetical 201A-derived powerplant.

We now have a choice: we can bolt a plate between the two upper ends of the engines (keeping the manifold geometry and attachment sealing a bit more straightforward; this would probably be better for engines cast en bloc which the 201A isn't), or we can make individual units with required fire rings, ports, and injectors for pairs of the power assemblies (comparable to a double-sided 201A head with circumferential instead of poppet valve-in-head exhaust and lateral instead of inline injection) which would be easier to work with.

Technically there is a third alternative, and here is where the way I see the semantics may be clearer: you can make up the equivalent of two opposed power assembly cylinders with the porting and injector bungs, etc. incorporated, perhaps by welding,  and then honed as necessary to seal pistons without scuffing as the engine heats in service (analogous to that Stumpf uniflow).  Now to the extent this retains separate liners (they would have to be extracted from the bottom) you could still consider this 'two cylinders' in terms of some of the parts and maintenance, but the outer part of such a power assembly, and its function in normal operation, is still one cylinder, with one injection location's worth of pump tuning or whatever, and it runs and could be considered exactly as one long cylinder with two opposed pistons working in it, and could rather easily be phased 12 degrees, etc. just as in the FM engine's single long liners and cylinders.

Glad you liked my 16 44 reference....the 24 66 had 12 cylinders ... coincidence? I think not.

No, it is not... but it reflects the common adding of 200hp per cylinder bore, added inline to longer and longer engines... no mystery there at all, unless you have some idiopathic aversion to multiplying or dividing by two.

As for cylinder count I was referring to the way the double acting engine functions   as a four cylinder would. Four expansion strokes per revolution of the crank.

All you have to do is what engineers do, count the number of actual power adders (in this case, piston assemblies) that are actually doing the stroking.  (This will serve you in oh-so-good stead if we get into things like Quasiturbines, liquid-piston, and free-piston designs any time later...)

If there are only two cylinders how come it needs four exhaust ports?

Now I know there has to be some English thinking going on; this is like the rigged decision to adopt vacuum over air braking "because the vacuum test train stopped quicker".  The exhaust port design in the 38 ⅜ is designed to produce adequate flow while retaining other characteristics; it could as easily have had more or fewer than 'two ports' to attain that.  There is neither a requirement nor a conspiracy to have 'one port per 100hp' in the FM engine...

The FM design has only 8 power strokes.....but 16 power inputs to the final drive per crank revolution (two cranks geared together rotating as one crank) 

That is not how the FM engine is designed, nor (if you look at the proportions and detail design of the bevel gearing) how the crank torques result in final drive to the generator.  Most of the upper-crank torque from the 38 ⅜ is used to drive the blower, with the bevel shafting much more important for long-term congugation precision than for 'compounding' torque through the lower crank to the traction generator (which is driven entirely off the end of the lower crank).  The engine does not work like a typical 'U' engine (where two equal block and crank arrangements drive balanced on a single output shaft, as in the Bugatti WWI 16 aircraft motor).

It is not that hard to find either accounts of the design of this engine or the practical implications of some of its design details on the Net ... and you'll have some interesting reading in the meantime.

[Overmod]

Lastspikemike

I just know you have a sense of humour. Be assured that I do also.

I am well aware or would not be responding at these lengths...

At the end of the day, it IS just a sandwich.

 

[Overmod]

Lastspikemike

There is a bit of a catalytic effect in that posts I read but don't understand sometimes become comprehensible when I read another post some time later. 

Remember there is nothing particularly mysterious or magical about steam power (decidedly unlike British electrical magic smoke!  - I am still searching for the safety data sheet on that Lucas part number...) so if you have trouble it's because I didn't translate it into English well enough.  The problem then becomes that for me to make it clearer I'll probably have to use even more words, use more references, etc. as the more disinterested eyes glaze over some more.

Ed almost certainly has that page of Angus Sinclair's where he rips crackpot patent drawings with arrows that show the steam flow ... but nobody told the steam it had to follow, and so in practice it doesn't.  Quite a bit of common-knowledge steam "knowledge" is in that sort of category, and more... including quite a bit from Porta, who really knew better... does not work as 'advertised' for similar reasons (the intricately, arcanely calculated Lempor that never seems to work 'as designed' in practice being one that comes to mind).

There is a bit, in passing, on steam chests and cylinder filling (and several nested reference lists!) in this little downloadable article:

http://static1.squarespace.com/static/55e5ef3fe4b0d3b9ddaa5954/t/55e6373fe4b04afd122b821d/1441150783767/%23+DOMS-1_Chapelon.pdf  

You may enjoy reading some of the others in that 'white paper' series.

[gregc]

Overmod

There is a bit, in passing, on steam chests and cylinder filling (and several nested reference lists!) in this little downloadable article:

The Development of Modern Steam 1: Andre Chapelon and his Steam Locomotives

i believe the indicator diagram in the linked article is misleading to suggest steam at boiler pressure typically enters the cylinder.   The diagram below more accurately illustates that steam entering the cylinder from the steam chest is lower than boiler pressure.

as both diagrams illustrate, pressure drops during admission (the Chapelon diagram shows a very significant drop).   this drop illustrates the steam chest pressure fluctuations mentioned in this thread.

while the steam chest pressure will be at boiler pressure at very low rpm, i believe the difference is greater at higher speeds as flow needs to match consumption by the cylinders and a greater pressure difference is required (i.e. flow is proportional to pressure difference). 

consumption is of course reduced as cutoff is reduced and steam chest pressure may be higher

[Overmod]

This is of course germane to the original posted question.

It helps to remember what an indicator diagram represents (and, a little more esoterically, why we use integration to recover information from an indicator diagram like the one modeled in the image gregc provided).  In the original indicators, calibrated manometers were used; in more modern versions, fast-acting pressure transducers -- these have some lag and overshoot concerns but of relatively small concern for most work.  At any instant the reading from such a device shows the instantaneous pressure; it has been trivial for a long time to implement an 'opposite of a stroboscope' to capture very short-duration readings coordinated with physical excursion or position of parts of the engine (e.g. the valve rod or engine crosshead, or with a little more sophistication the physical or effective steam edge of the valve) and observe pressure deviation over time under different conditions.  Only incidentally are these readings 'integrated' by the indicator's charting engine into a 'continuous' trace or line.

It is pretty obvious that without some careful or exotic remediation the 'spot pressure' as admission just begins will drop, as mass in the chest begins to be 'accelerated' into the opening port faster than gas kinetics can support.  (In the admission diagram this is reflected by the little pink wedge 'b', although it need not, and probably should not, be progressively lower through the full duration of admission as drawn).

Having more hot steam in a 'properly sized and shaped' chest adjacent to the valve minimizes this drop, but will not eliminate it.      Fortunately perhaps the effect on engine performance is relatively slight, because the effective mechanical advantage of piston thrust in this range -- depending of course on the point of 'limited cutoff' less the effect of any slot/Weiss porting, Herdner valves or the like -- will be limited up to the point flow down the port into the passage and cylinder volume 'equalizes' with the flow into the chest from the tract so that EP falls little with admission mass flow as great as needed to fill the swept volume (behind the mechanically-accelerating piston face).

The thrust concern of the drop here is relatively small compared to the mass-transfer consideration.  We have not yet looked at the individual pressure trace at the valve in the instants after cutoff (where flow from "the tract" into the port, which was ostensibly flow-smoothed, 'streamlined', assisted by steam conditions in the chest, etc. in passage through the chest, is quickly interrupted in the last moments of crossing the port edges (see the 'trapezoidal port' issues in general for more) but as mentioned the 'practical' effect here is of no particular interest to the engine designer unless any deleterious effects persist until the instant of admission at the other port -- and that time is an eternity compared to the recovery time over admission-flow establishment and then actual admission flow at any reasonably short cutoff.  The pressure drop will have recovered and any oscillations or shocks 'ringing' in the chest volume will have either been damped or evenly distributed in steam kinetics by the time of next admission, and any attempt at 'tuning' to achieve enhanced early mass transfer a la cross-ram induction is likely to be difficult across the required range of speed vs. mass flow in this size and cycle time range.

I would be tempted to consider this as if it were a structural loss like nucleate condensation, because many approaches that might reduce it will either impact the cost of the engine or its flexibility in operation -- or its maintenance.  Some of the prospective "solutions" introduce the same potential for catastrophic consequence upon derangement or failure posed by various kinds of nonproportionally-linked servo-driven valve gear, a 'design improvement' that armchair thermodynamicists perennially propose as if it were their bold new insight into a better reciprocating locomotive.

Most of the rest of the comments are right, but the concern is not that they 'exist' but what can be done about their effect.  The 0.85 factor in the classical formula reflects these much more than putative 'heat loss' from boiler to valve that some people seem to assume.

[gregc]

would you mind commenting on the difference in steam chest pressure from boiler pressure and the reason for that difference at different rpm?

[gregc]

at some point, with wide open throttle and significant cutoff, steam chest pressure will be insuficient to go any faster pulling a train on level ground.

gregc

how close to boiler pressure is steam chest pressure or MEP cylinder pressure at max speed on level ground?

can this be quantified if we know the rolling resistance at that speed (see Armstrong's charts)?

i believe yes

[Overmod]

Lastspikemike

Those cylinder heating steam jackets would reduce pressure loss on transit of steam from steam chest to cylinder.

It was my impression that active jacketing (either with exhaust steam or overcritical tracers) does comparatively little in reducing passage pressure loss -- their contribution is to provide lower heat loss in the passages, and net of acquired compression heating this is not much improved beyond the first few sustained strokes.  (Remember there is still full thermal cycling of the 'contact' face of the passages over that ~.007 range of depth, meaning that contact of the steam during the comparatively brief admission is not going to improve over the range of cycling due simply to the inherent steam heat content between admission and exhaust...

Wardale in fact makes a fairly strong case for nothing more than good lagging and insulation around that area, that anything added by extra jacketing (and its complications) or tracers doesn't justify the trouble and expense.  (He uses steam cooling of the valve liners to solve tribology concerns there due to high inherent superheat, some of which may help wire-drawing condensation concerns as the ports open to steam in gregc's period of interest...)

There are some interesting Google Books treatises on steam-engine jacketing (in smaller road-engine practice, where it becomes more essential and the bulk involved less loading-gage critical) which you might want to read critically and assess.

Pre heating the Newcomen cylinder would have been even more efficient, by exhausting the still hot steam into a cylinder jacket first before condensing it.

If you work through it, you will see this really doesn't help anything with cycle efficiency: the idea is to get the heat out of the steam mass ASAP, and the added flow impediments involved in passing that steam through a jacket (remember that the cylinder inside wall was at 'steam temperature' before, and your exhaust jacket temperature is by definition much lower) will impair doing so.  

Likelier better is to recover heat in the exhaust steam somewhere in the condenser, if possible, and use that in the Rankine cycle somewhere (probably as combustion-air or feedwater heat).  Historically though the full 'sinking' of that recovered heat is either uneconomical or saturates the available means long before high sustained engine power via mass flow is demanded.

In a Newcomen engine, one of the early 'improvements' was better lagging/insulation of the outside of the cylinder, and it would probably make sense to put one of the stages of the "condenser" on the outside of this lagging, to 'thermally float' the acting cylinder as much as possible and to cut down heat loss outside the lagging (ideally to negative flow, but that would not likely balance at the inner wall of the cylinder during air exclusion, which in a Newcomen engine is what you're optimizing).

I am a bit chuffed to have understood that enlarging the steam chest to match that of the cylinder to be filled improved efficiency.

You should be.  There are still many 'experts' who really ought to know better who still dispute the value of the idea; it can in fact not work right if all the sorts of refinements Wardale apply to it are not present.

The idea of reversing the "low" and "high" pressure cylinders in terms of flow was particularly interesting, in effect superheating twice to lower maximum temperatures but extracting and employing at least  the same amount of heat from the fuel.

I assume you're talking about the 'original' of Chapelon's 2-12-0 here, which fed high-temperature saturated steam around flow-optimized passages in the HP cylinder block before admission, then used Schmidt-type 'superheater' elements for reheat (including superheat) going to the LP cylinders.  This was interesting, but I believe he discarded using the arrangement during testing; see some of the discussion Thierry Stora and Claude Bersano provided about it.  (Or contact Juniatha via PM as she has carefully analyzed the arrangement.)  

This is a sort of flow-optimized version of a 'large steaqm-chest volume' approach, and if you look carefully you can see that the effect of superheat present in the admission steam (there is very little by design in Chapelon's arrangement as built) is thought to contribute little to gregc's admission pressure drop of concern.

The exhaust Venturi stacks were a clever application of the same idea utilized for injectors  feeding water from the tender to the boiler.

It's actually a different application of the same set of physical principles: the injector trades velocity for pressure, whereas the 'vanes' provide more divided precise flow shaping and vector matching of combustion-gas flow going to entrainment in the exhaust jet -- my opinion being that optimizing this physical entrainment is an important physical design principle in high-speed front-end design, something that not all the steam engineers give the same weight.  The Kylchap applied to the fastest A4s is notable not only for its refined stages of flow entrainment, but the absolute length and feed distribution of the array.  

Here it pays to study the Master Mechanic front end and its principles, which are very different from most British boilers.  Carefully chart the flow path through one of these things, making careful note where the baffle plates go and what they do -- one of the effective purposes is to constitute a convergent-divergent nozzle arrangement on a grand scale in the combustion-gas flow, incidentally greatly increasing the actual gas-path length ahead of the front tubesheet.  You'll see much mention of the mysterious 'lip' that was apparently so sensitive in early empirical front-end tuning; I doubt you will have any trouble winkling out both what its purpose was and why it is such a sensitive thing to tune -- and yes, you should be even more chuffed after realizing those things.

Makes you wonder how far the steam locomotive might have progressed had it not been for the nuclear power plant, which converts steam energy to electricity...hmmm mini nuclear plant powered locomotive anyone? Postwar, France invested heavily in nuclear powerplant generation which made the TGV feasible if not economically sensible.

Interestingly, all the action with murdering advanced steam-locomotive practice in favor of electrification occurred well before France had a clue about nuclear steam generation, let alone built it out with enough density to power 22,000-odd horsepower trains on a regular basis.  The use of nuclear for railroad principles only makes sense by using commercial-scale nuclear grid stations; the most distant runner-up is nuclear batteries, and the security concerns alone for those are generally appalling.

Incidentally the thermodynamics of nuclear stations are important -- they are wretched even in comparison to traditional locomotive practice, only made practical by lavish but very cheap heat flux.  (The same thing that makes nuclear blue hydrogen an attractive prospect ... but I digress.)  While it is possible to superheat nuclear steam, it is very difficult to do so safely (the only Russian approach would curl your hair to see, and at the original Indian Point steam station it was projected to do it with separate oil-fired superheat as in marine practice) so all the PWR/BWR designs in the 'golden age' of power construction ran on saturated steam (admittedly at very high throttle pressure) and this is a major contributory reason why nuclear plants "can't" be rebuilt with renewable or conventional steam generation and run cost-effectively.  The advanced gas-cooled reactors can have less limitation in this respect, but it remains to be seen if those can make a return to large markets in an age when 'zero-carbon' becomes essential but FUD regarding anything with 'nuclear' associated with it anywhere is still strong and amply advocated...

[Overmod]

gregc

at some point, with wide open throttle and significant cutoff, steam chest pressure will be insuficient to go any faster pulling a train on level ground.

And very likely it is the combination of pressure drop and effective flow establishment near the port area in the steam chest that constitutes the limiting factor on "top speed" in a piston-valve reciprocating locomotive, where the amount of cutoff needed to sustain any additional acceleration or speedholding -- the aforementioned 41-45% range observed for a large number of good piston-valved designs -- gives the point where mass flow from the boiler just balances what makes it into the cylinders.  Very likely, the 'next' factor that would limit this would be "inadequate" steam-chest reserve close to the ports, although at the high cyclic involved I suspect actual intermittent flow-interruption effects would become minimal and in fact actually lead to the possibility of a little enhanced early-admission pulse-tuning effect.  In practice the greatest possible improvement that could be made in this area would produce very little practical speed increase, at most a mph or two, and that principally by relieving traditional mass-flow losses at high cyclic for reasons other than flow establishment at the admission porting.

Very likely mistakes here are the 'valve problems' observed in some otherwise-fairly-competent-looking designs that ought to have been able to reach very high speeds but most certainly didn't -- the ATSF 3460 class perhaps being the poster child of all poster children, but the C&NW E-4b a very close (and prettier sister!) runner-up.  The 3460s were easily capable of speeds right up to the 100mph sustained range, but the performance then fell off a cliff, with great difficulty getting to 102-3, and the peak achieved being about 105.  This very clearly suggests ports as being the trouble spot ... possibly the only limiting restriction on getting to T1-like levels of sustainable speed, but in the event a sufficient limiting restriction.  The North Western certainly didn't lack for good high-speed designers, but AAR couldn't get those Hudsons more than about 7mph faster than fairly stock PRR K4s(!) which are nowhere near the design an E-4b otherwise is.