Dear Pete,
Thanks for the clarification- a treasure trove that I need to get my hands on. Particularly interested in the De Glehn, as it ties two continents' data together.
Dear Paul,
Thanks for these comments and ideas.
With respect to fluctuations in cylinder temperatures and the reason for the negative impact of condensation on LP efficiency, let me give you my understanding of the situation.
There’s no problem with saturated steam being formed in a cylinder during expansion, as you suggest. However, I’m talking about condensation not of steam in the cylinder, but of the steam onto cylinder walls, something which does impact efficiency. If the cylinder wall temperature is below the saturation temperature of the inlet steam, then a film of water forms very rapidly on the cylinder surface at admission. Since water is a poor conductor, it very quickly reaches a limiting thickness - up to 0.1mm. (Thickness depends on P1/2, and the cylinder temperature). Now during expansion, the steam in the cylinder falls below the temperature of the cylinder wall and its film, and the water film re-evaporates, and disappears up the exhaust without having done any useful work- a kind of leakage. The amount lost also depends on the surface are of the cylinder; on smallish UK cylinders, the amount can reach 1000lbs/cylinder/hr, not a large amount, but sufficient to take the edge of high efficiencies. I’m maybe making the LP losses through condensation more dramatic than they likely are, but they could eat away several % of any expansion ratio benefit compounding brings.
Question then is, what is, and what determines cylinder wall temperatures. These questions were addressed at Altoona in 1912; obviously the temperature lies between inlet and exhaust temperatures, and depends on the temperature of the inlet steam, and cut off- the greater the exposure to live steam, the higher the temperature. Knitting these data together with a theoretical model which calculates the size of the film under various operating conditions allows you estimate how much steam is wasted by this process. It is this kind of analysis that says that, in the HP cylinders of Compounds, where the exit temperature is always higher than the saturation temperature of the inlet steam, there will be no wall condensation. In the LP cylinders, where wall temperatures will be lower, there will be, despite the lower saturation temperature of the inlet steam.
This condensation process is the main source of heat transfer losses in cylinders. The Altoona data do not mention temperature fluctuation during the piston cycle, and Professor Hall, who developed the quantitative cylinder condensation model mentioned above built an apparatus to measure cylinder wall temperatures over the cycle. What he showed was that the block as a whole reached an equilibrium temperature, and if no condensation occurred, as with high superheat, there was no fluctuation in the temperature of the wall. When condensation did occur, there was a fluctuation in the very top of the surface, as heat flowed in from the condensation, then flowed out as the water film re-evaporated.
http://5at.co.uk/uploads/Bill%20Hall%20software%20and%20papers/The%20Effect%20of%20Superheat%20on%20Cylinder%20Condensation.pdf
So, clearly, the condensation process apart, there is little heat transfer to and from the cylinder walls during the piston cycle, once the equilibrium temperature of the cylinder block has been reached. Chapelon feared heat losses along these lines, but his fears were I think unfounded.
As to what would happen to cylinder wall temperatures if you went to very high superheat, obviously there’s no data, but like you I suspect any problems are surmountable. However, there is another effect worth considering which is that as you increase superheat, so the temperature of the exhaust goes up, and with it the back pressure at a given exhaust flow. With conventional exhaust systems, this means that once inlet steam gets up to 8000F, you are already getting into diminishing returns, and by 1200oF the benefits disappear and things get worse, according to my calculations . I suppose you would be able to optimise the exhaust to deal with this, haven’t thought about this though.
With respect to your comments on lessons from modern power plants, perhaps the biggest lesson is that if you want to use steam efficiently in a heat engine, turbines are the way to go? Come back Jawn Henry? Seems to me you would only go down this road if there were only coal left to burn, and the economics of electrification didn’t add up.
Dear Neil,
Thanks for this info from Giesl. HIs conclusion about typical steam rates with NYC Hudsons conflicts with my own analyses, which suggest that no more than 3500IHP was required to gain significant amount s of time of the fastest schedule, the 20th Century with about 1000 (US) tons, no more than 50000lbs/hr cylinder flow, say 45000lbs/hr blastpipe flow with 10% feedwater recycle. I seem to have found some data that the J1s blastpipe as having increased from 6.75" to 7 3/8" over time, but with 6 sqins lost due to 1/2" bars. At the lower area, you would indeed get blastpipe pressures around 1 atmosphere, at that kind of steam flow.
You will find me sceptical to the point of tedium on most claims from the steam age (maximum speeds, powers, efficiencies etc) especially where there is a personal or company interest involved. Many don't stand up to detailed scrutiny, or are data massaged in such a way as to present the interest group in the best possible light. (Nothing new here, then). So, I feel like a bit of a curmudgeon, but getting to the bottom of matters is what interests me.
I was interested in your comments about maximum German evaporation rates. I haven't looked at German practice much, but I did do some analyses of running at the end of steam from Bremen to Osnabruck, and ws surprised to find that the running of a heavy 600tonne train which gained 10 minutes on the schedule was done with little more than 2000IHP, this from an oil fired Pacific ( no fireman limit) with a 43 sqft grate, no more than about 600lbs/sqft/hr depending on boiler feed system Given this effort was significantly in excess of what was normally needed, I formed the impression that German boilers were not steamed hard.
I corresponded extensively with John Knowles during my researches into Locomotive resistance, and the basic approach we adopt is the same, seeing overall locomotive resistance as the sum of its resistance as a leading vehicle, given by the Davis equation, and the resistance of the machinery between the pistons and the wheel rims (MR). This, I later discovered is the Baldwin/RP Johnson approach, much more realsitic then the German Strahl equation. We were actually quite close in our estimates. Where we differ is in the value of MR; he uses an empirical approach which gives values somewhat higher than given by the first principles anaylsis done by a mutual friend mentioned in another post. I think these first principles figures are in much better agreement with UK test plant data for MR, and overall Locomotive resistance as measured in road tests- but we are talking relatively minor differences on a topic where there are very widely divergent published views, that do not really affect the outcomes mentioned here. Interestingly, the value of MR proposed by Johnson is I think way too high. It is based on Altoona testing, but these figures can be shown to be erroneous, a minor blot on an otherwise priceless set of data.
Paul Milenkovic The real limit on superheat temperature is the superheater tubes. Otto and Diesel cycles can have very high cycle temperatures because the applied heat is intermittent, and there is also water or air cooling on the outside of the cylinders and cylinder heads. Gas turbines cannot have quite as high turbine inlet temps, but again, there is a hot side and a not-hot side of the turbine blade, and various cooling schemes can be applied. With a steam superheater, the material has to withstand the full superheat temperature on both sides of the tube, which I see as a limit to steam cycle temps, even if you go to exotic "superalloys" for the tubes as you do for turbine blades.
The real limit on superheat temperature is the superheater tubes. Otto and Diesel cycles can have very high cycle temperatures because the applied heat is intermittent, and there is also water or air cooling on the outside of the cylinders and cylinder heads. Gas turbines cannot have quite as high turbine inlet temps, but again, there is a hot side and a not-hot side of the turbine blade, and various cooling schemes can be applied.
With a steam superheater, the material has to withstand the full superheat temperature on both sides of the tube, which I see as a limit to steam cycle temps, even if you go to exotic "superalloys" for the tubes as you do for turbine blades.
Paul,
My understanding is that the amount of superheat is limited by water disassociating into OH- and H+ at much higher rates than at room temperature. Both ions can be very corrosive and make life miserable for any metal in their presence.
- Erik
One of the better early compounds was the Bavarian 3/6 built between 1908 to 1931. Even in the 1950s they were preferred by many crews to the larger 01s and 03s on the hilly curved route from Munich to Lindau. Coal consumption for a round trip was about 2 tonnes less and oil consumption was similar despite the newer engines having only two cylinders. Their better balancing allowed them to start more easily than a two cylinder engine and acceleration up to 62 mph was better. Above this their machine resistance and corresponding wear rose rapidly. Maximum speed permitted was 74 mph (coupled wheel diameter was 1870 mm). Compared with the 3 cylinder 01 and 03 10s they were easier to lubricate (all had bar frames). The workshop personnel preferred the simpler 01s and 03s but overhaul costs were of the same order because damage to frames was unusual in an S3/6 on account of their smaller and better balanced piston forces.
Minimum coal consumption was 5.8 kg/ IPS, an excellent figure corresponding to about 795 C steam temperature in a simple expansion engine. With tube length limited to 17.25 ft evaporation rates up to at least 70 kg/(m2 h) were attained (2300 IPS), well over the official 1770 to 1830 IPS.
There was a small series with 2000 mm coupled wheels, which were used also outside Bavaria, including on the Rheingold in the 1920s and 1930s.
Despite their modest adhesive weight of 52.7 tonnes they handled trains of over 600 t. 18 478 is again operational.
They were a well balanced type with strong frames and a good boiler, as well as the good balance of their 4 cylinder layout.These other factors make it is hard to say how much their being compounds contributed to their success.
Thank you for your reply. I must correct an error in my post though - only 6 compounds.
The 8 tested were:
1. simple PRR 2-8-0 type H6a
2. simple LS&MSR Brooks 2-8-0 class B-1
3. cross compound MCR Alco 2-8-0 class W
4. tandem compound AT&SF Baldwin 2-10-2 class 900
5. DeGlehn PRR SA de CM essentially an exact duplicate of those supplied to the Northern Rlwy of France
6. Vauclain balanced compound AT&SF Baldwin 4-4-2 class 507
7.balanced compound w/ superheat Royal Prussian Rlwy Admin HMAG 4-4-2 class S8
8. balanced compound NYC&HRR Alco 4-4-2 class I1
Pete
I think we are pretty much in agreement that steam locomotive 2-expansion piston compounds can have losses between the HP exhaust to the receiver and the LP intake from the receiver. As I mentioned earlier, this shows up in compound indicator diagrams where there is the HP pressure-volume loop, the LP pressure-volume loop, and this big gap between the two loops representing what you call entropy loss (i.e. lowering in pressure without doing any mechanical work).
With respect to the thermal losses I had suggested, one of the knocks on steam-locomotive piston engines is that you have alternating hot inlet steam and relatively cooler exhaust steam flowing in and out of the the ends of the pistons, and such things as the Skinner Uniflow engines, used on the Lake Michigan ferry boats but never in railroad applications are supposed to reduce thermal losses from this effect. Separate into HP and LP cylinders, in the words of a gimmick sandwich package from a fast-food restaurant, are supposed to be better at "keeping the hot side hot, the cold side cold." What I am thinking of are conductive losses, of heating a chunk of metal in the cylinder with inlet steam and then cooling that same chunk of metal with exhaust steam.
With respect to conductive losses, why is condensation, especially in the LP cylinder so bad? I can understand you don't want too much fluid in the piston that you burst something with hydro-lock, but why the worry about saturated steam? The only thing I can think of is that if steam is condensing on the cylinder walls, you are transfering enormous amounts of heat as losses to those walls whereas if the steam stays as steam, you get much less heat transfer to the walls?
With respect to high superheat temperature, I don't think that the upper limit on superheat should be lubrication of the cylinder. You probably don't want high superheat during "hard working" -- 85% cutoff -- but with respect to high superheat under 15% cutoff where you are getting a lot of expansion, the high piston temperatures are only intermittent, much as a Diesel engine that they know to keep oiled in the face of very high intermittent temperatures. On the other hand, oil is typically injected into the steam, which has the effect of exposing the oil to full superheat temperature even though the steam cools down upon expansion in the cylinders. Perhaps one would have to inject lube into the cylinders more directly.
In light of this, the most efficient steam cycle, within the limits of boiler pressure brought on by scale formation in a cycle that is not using highly distilled water, and within the limits of not using a condenser as you don't have access to river water in a moving steam engine and you do in a stationary power plant, that steam cycle should use a high degree of reheat between HP and LP stages, yeah, yeah, the reheat tubes and headers also add to pressure drop.
I am thinking that not only do you want the exhaust to be above condensing, you want the exhaust to approach the boiler temperature. That way you could use exhaust steam to not only boost the feedwater temp up to atmospheric boiling of 212F, you could boost the feedwater temp up towards the boiler water temperature of around 400 F? One would have to work all the details of the thermodynamics, but I believe highly-optimized stationary powerplants use reheat between compound stages along with transfer of heat from LP exhaust to boost the feedwater to a high temperature (the feedwater heater needs to apply heat to the feedwater after injection to boiler pressure for this to work). And if automagically you could use this excess exhaust heat to pre-heat the combustion air, so much the better.
Before you think I am crazy, the British 5AT project (they have a Web site) has some kind of metal partition and use of the front part of the boiler at the smoke-box end as a kind of preheat-the-injected-water-to-boiler-temperature setup, but they are using the leftover heat from combustion in the flues just prior to the smokebox. I am proposing using exhaust steam heat for this purpose by applying reheat in a compound to boost exhaust steam temps, and I am thinking the power plants already do this.
But yeah, yeah, in the steam locomotive application, these schemes to wring out more efficiency involved more hardware (I am not giving up on compound expansion and I am adding a reheater in addition to the superheater along with a high-pressure feedwater heater), and in the end, it was the maintenance costs not the fuel costs that were the dominant economic effect.
If GM "killed the electric car", what am I doing standing next to an EV-1, a half a block from the WSOR tracks?
The $64K of question, of course! I drafted my thoughts on French Compounding into an article for a general audience, deliberately skating over some of the hard questions. I circulated to local experts who all came back and said that the fundamental questions needed tackling head on, so a complete rewrite needed. I now have four pages that attempt to answer your question, enough to send even the most dedicated reader to sleep. So I have had a go at trying to simplify it in response- still over a page- so see what you think. (I'll get to your specifics at the end).
Why Compounding?
There are two ideas to consider.
Firstly, improve expansion ratio, the volume of steam in the cylinder when exhaust ports open divided by volume of steam in cylinder at cut off, including clearance volume. The higher this is the better, providing the expansion is isentropic, as it is reasonable to assume in a simple engine. If you calculate expansion ratio for a compound, even flat out in 70% cut off in both cylinders, and allowing for 25% CV in the HP cylinders, this is in excess of 3, whereas for simples, the range varies from about 1.5 to 2.5 as cut off reduced from 50% to 20%. There is however a problem, not always recognised in steam age literature, namely that the expansion from HP inlet to LP exhaust in a compound is not isentropic, because of what happens between HP exhaust and LP cut off, so the above ratios are misleading, and the expansion ratio benefit is not actually known from these simple sums.
So the second, underlying idea is to maximise isentropic efficiency, in other words reduce irreversible losses associated with increases in entropy. The maximum amount of work possible from a heat engine-the isentropic (constant entropy) ideal is, for a given inlet temperature and pressure and exhaust pressure, fixed by the second law of thermodynamics. There are in fact four major sources of irreversible losses:
So how do Compounds fare on each of these criteria?
What’s important to recognise in this is that if you take a simple engine working in short cut off, and high superheat, i.e. no condensation, at realistic backpressures you can in principle get isentropic efficiencies of ca 85%- the cylinders are doing a pretty good job of getting the most that is possible from a given inlet temperature and pressure. The ‘in principle’ refers to the fact that, from UK test data, there is usually at least 5% leakage to take into account, which means that isentropic efficiencies in the low 80s were in practice the best achieved.
So, if a good simple can get 85% isentropic efficiency, there’s only 15% to go for, whether through compounding or valve gear design. The Chapelon 4-8-0s did, by my as yet provisional calculations achieve isentropic efficiencies in the low 90s, 10% better than actual simples. However this figure only applies if a) the superheat is high enough into eliminate LP condensation, and in most other cases it wasn’t and b) the right combination of LP and HP cut offs is used, which depends on both the design of the engine and the actions of the crew and c) there is no pressure drop between HP exhaust and LP inlet.
The design constraints on a steam locomotive are very considerable, which means, I think, that it is difficult to get the few % improvement in thermodynamic efficiency that is possible, which is why most abandoned the attempts to do so. Compounds were successful in a marine environment. The Titanic had triple expansion Compound engines rated at 46000HP. However, there are no hills or stations between Southampton Water and the Hudson, nor could the operating department add extra coaches for peak periods, nor was adhesion an issue, nor were there constraints on cylinder size- a much simpler environment.
With respect to you specifics;
a) It's clearly an objective in any engine to reduce wiredrawing, hence attempts to improve valve size etc. Whether, because Compounds were working at high HP cut offs against a higher backpressure, hence with higher pressure at cut off there was a net benefit is hard to say, but bear in mind if there's only 15% to go at, and a lot of this is due to incomplete expansion, it's maybe not as big a deal as people thought. Also bear in mind that you will have a second set of wiredrawing losses in the LP cylinders.
b) Not sure what you mean by 'reduced thermal losses'- do you mean e.g. convective and radiative losses? These seem to be small
c) Cylinder friction losses are not that high. I have worked with a friend to produce a good quantitative model of Locomotive resistance. A key element of this is to estimate 'machine friction' losses from first principles, which includes of course losses associated with cylinder friction. The friend's estimate comes out as rising from 3HP at 20mph to 14 HP at 90mph for a large UK 4 cylinder Pacific. Even if this is out by a factor of two, it's not such a big deal.
d) On the earlier French Compounds, the LP inlet pressure was below the HP exhaust pressure, so there was an unhelpful pressure drop of the kind you indicate (see above). Chapelon's development work elimiated this- so it is possible with the right valving.
Thanks- most interesting- more evidence that early 20th Centrury Compounds weren't fulfilling their potential. I did read Warden's book on the NW Ys in the unfulfilled hope of getting some test and technical performance data on them. He clearly was a big fan of the Ys, but never talks about efficiency benefits, and himself wonders if it would not have been better for the NW to have built more versatile As instead of the last Ys.
What is the compound supposed to do?
Is one thing to get a greater expansion ratio of the steam within the limits of a particular type of valve gear (i.e. Walschaerts)? That is, if you had a simple with the right valve gear without the "wired drawing" (throttling) limitation at short cutoffs, it could achieve higher expansion ratios?
Is another thing thermal, that by expanding in separate stages you keep the HP stage hot, the LP stage not quite as hot, and reduced thermal losses?
Is a third, possible one of cylinder friction, that cylinder optimized for a certain inlet pressure you have a certain amount of friction that there is a diminishing return for short cutoff and high expansion ratio? That a LP cylinder can have lower friction than a HP cylinder of the same volume on account of the lower steam pressure to seal against.
Working against any efficiency gains of compounds appears to be indicator diagrams I have seen. It seems there is considerable white space on the indicator diagram separating the HP indicator loop from the LP indicator loop, suggesting there are losses in the HP exhaust, LP intake and the receiver in between to contend with. Does this make sense, that compounds incur losses in transfering steam from the HP to the LP cylinders?
Dear DreyfussHudson,
Thank you for your interesting reply. I agree that some historical data seem inconsistent or wrong, even from Swindon and Rugby. For an interesting but very demanding review of locomotive resistance, people might be interested in a piece 'Steam Locomotive Resistance' by John Knowles that has been posted on the web.
Regarding the West Country, a steam rate of 25000 lb/h seems typical of everyday use. Looking at curves of blastpipe pressure against steam flow rate for various degrees of superheat given by Giesel, for a pressure of 0.13 atu, flow rate is around 40 kg/(cm2 h). For 25000 lb/h this needs an area of 283 cm2 or 44 in2 (7.5 in diameter single blastpipe). This seems too high given that the Kylchap in the A4 was 39.23 in2. The only example I have found so far for a Giesel is for a Class 78. This was a medium power fast tank engine. When new in the 1930s, Giesel says blastpipe area was 128 cm2 but by the early 1950s some had only 95 cm2. Perhaps they had steaming problems similar to some British types postwar. From 1956 they received Giesel ejectors and higher superheat (through throttling gas flow through the large tubes while easing it in the small tubes). Giesel gives the ejector area as 232 cm2. This corresponds to a single blastpipe of 6.75 in diameter. Nominal steaming rate was 12200 kg/h or 26900 lb/h. The 7.5 in blastpipe would only look consistent if Giesel thought the nominal steaming rate of a West Country was 32800 lb/h with an all-out maximum of at least 36000 lb/h.
Regarding his time in the US he says he rode on a NYC 4-6-4 during tests in the early 1930s. The engine was one of the 1927/30 batch. With coal was of 7000 calories he says and normal evaporation rate was 60000 lb/h, maximum 70000 lb/h. This corresponded to an evaporation rate per sq m of heating surface of 77 kg/(m2 h), 90kg/(m2 h) maximum. This he says was far over the prewar German limit of 57 kg/(m2 h) and even their 1950s 'high performance boilers', which were 75 kg/(m2 h). Blastpipe pressures were a high 1.2 resp. 1.6 atu. He says the exhaust sounded like gunfire. He comments that the gas free area was relatively good. Further, in connection with the high specific and especially absolute boiler tube resistance, together with the self-cleaning smokebox, and the high performance, the overall design was good. Certainly better than the German.
Dreyfusshudson ...All of which says that when experiments were being done on Compounding in the early 1900s, it is quite possible that there would be little or no efficiency benefit for Compounding vs a well optimised simple.
This is an interesting statement. The B&O EL Class were originally built around 1916-1920 as compound articulated locomotives. In the late 1920's into the 1930's the B&O did a rebuild of the ELs in their own shops, converting them into simple articulated locomotives. Among the many changes, the boilers were redesigned to increase the draft and increase the superheat temperature. The test results not only showed a large increase in horsepower at speed, but an overall increase in efficiency. The final result was they took a locomotive that was designed as a slow speed drag locomotive, and ended up with a locomotive that was quite good at pulling manifest freight at moderate speeds, with a decided HP advantage over their S Class 2-10-2 (which were highly regarded freight haulers in their own right). The ELs ended up in service on the B&O right up into the mid 1950's before being scrapped.
Many thanks- I was unaware of the references you quote, and they are of great interest. Having completed the ‘American Project’ and written it up into an article for publication, I turned my attention to the feats of Chapelon’s Compounds, which in the 1930s were shown to be able to deliver more than twice the power at the drawbar of contemporary simples of similar size in this country (UK). There are two possible kinds of explanation of this; firstly the Compounds were markedly more efficient, so produced more power per unit of coal fired; secondly, the enhanced draughting allowed evaporation rates/sqft grate that were much higher than achievable with the simple draughting devices used in the UK at that time. This has been an enormous endeavour, wading through hundreds of pages of French text, and simulating the behaviour of Compounds using the computer programme.
Apart from possibly the Chapelon 4-8-0 (there are apparently conflicting sets of data in the same report), there is no evidence that French Compounds had any significant boiler or engine efficiency advantage over simples with the same superheat and exhaust pressure. The French Compounds were way ahead of most UK designs in these two respects in the 1930s, so will have had economy benefits (Explanation 1), maybe 20%, meaning overall, explanation 2 is the more important one, and if you work out the evaporation rates/sqft required for the highest claimed outputs, they could only be attained with superior draughting. However, if you then look at the coal rates required to achieve these steam rates, they are generally way beyond the amount it is reasonable to expect any fireman to achieve (most Compounds were hand fired). So, whilst the drawbar power claims are factually correct, they are simply demonstrations of potential, not of what might be expected in service, which, at similar efficiency would be no more than a simple could achieve if fitted with the same draughting and raised superheat. Many of the cylinder powers claimed are several hundred HP too high- Chapelon had a very unrealistic view of Locomotive Resistance- hence cylinder efficiencies also.
The Computer modelling of Compounds suggests that, as has often been stated, Compounding can provide up to 10% benefit in efficiency. This depends on exactly what you choose as the simple to compare with the Compound. If you fed live steam to both the HP and LP cylinders of the Compound i.e. operated it as a simple, this would have similar engine efficiency to the Compound, but you would have a brute with enormous tractive effort relative to adhesion weight, which would have to be worked in extremely short cut offs at speed- in practice an non-starter. It is against practicable simples that the benefits begin to show.
However, a number of factors can eat away at the 10%. Some of these issues Chapelon rectified, meaning older Compounds would not achieve 10%, and 10% also requires that you use the right combination of HP and LP cut offs. Most importantly, my calculations suggest that unless the inlet steam temperature is above 750oF (as on the 4-8-0, but not the other Compounds), there will be significant condensation on the walls of the LP cylinders, reducing efficiency; on simples, cylinder wall condensation ceases above about 6300F.
All of which says that when experiments were being done on Compounding in the early 1900s, it is quite possible that there would be little or no efficiency benefit for Compounding vs a well optimised simple. The Great Western Railway bought four De Glehn Compounds in this period and came exactly to this conclusion. That the de Glehn system was actually tested on the PRR plant is very exciting news, for one should be able to analyse the results to find out exactly what was going on in machines of this era, which by and large sealed the fate of Compounding. If the De Glehn reports do not form part of the numbered sequence of Altoona reports, then they are probably not at the library in London. I have found a source in Google books, The Pennsylvania railroad system at the Louisiana purchase exposition: locomotive tests an ... but unfortunately this exceeds the 700 page limit, so will not reveal its contents. I may be able to source this in the UK.
Just in case the PRR test results for 8 compound engines tested in St Louis in 1904 come within your sphere of interest and are not in the 2 bound volumes, and if no-one has already mentioned this, here are details:
The test reports are not numbered, as with other test bulletin numbers but identified by title only, eg 'Tests of De Glehn Atlantic Type locomotive, PRR'.
The test results are in the 1905 publication 'The Pennsylvania Railroad System at the Louisiana Purchase Exposition'. The volume includes a detailed description of the Altoona test plant as erected in St Louis, together with chapters on calibration of instruments and methods of recording and conducting tests.
Cheap, print-to-order copies are available from ABEBOOKS. I was thrilled to pick up a 1905 copy in a PRR memorabilia fleamarket stall near Altoona before the advent of the web made such things so easy, albeit now at a price ($hundreds).
Regards,
Pete1950
This responds to your three recent posts.
Firstly, with respect to the inlet steam temperature on the Jubilee, I had it rising from 298 to 3260C as steam rate increases in line with the test data on 45722. Steam chest pressure was assumed to be 215psig. I should say that, after ten years of using this fluid dynamics engine programme, it is in my view in a different league to anything available in steam days, providing of course, you know what you’re doing! Giesl’s approach is now superfluous.
On the Giesl results, I’m somewhat sceptical of Giesl’s claim. In UK test plant data, there are plenty of examples of misleading backpressures, due to problems with where the backpressure was measured. The programme mentioned does compute blastpipe pressure, in line with standard discharge theory. If you look at most Test results from Rugby, the agreement is very good between experiment and theory. There are a number of case where, for a given free nozzle area, measured blastpipe pressures are higher, but this is on locomotives where an orifice plate was fitted to the blastpipe. (An American idea, I believe). This reduces the discharge coefficient from 0.99 to about 0.90, so the effective area (working back to a common 0.99 coefficient) is about 10% less- so not as good as you might believe from a backpressure point of view.
There are then locomotives which show lower back pressure than they ought. Amongst those tested at Swindon, there are instances where the reported pressure is about half what it ought to be (Duke, DC, King DC, MN Lemaitre, to a lesser extent King Single chimney). Indeed the MN results at Swindon are about half of what Rugby reported (and the Rugby results are slightly lower than you would expect). At Rugby, the Duchess DC is about half of what it should be, and half of what the LMS reported. I have worked through the Duchess case carefully, and the problem is due to the fact that the pressure was measured at the base of the blastpipe, and there is very slight narrowing to the orifice, hence pressure increase at the tip. I suspect, but haven’t proved that this is the problem at Swindon.
Rugby did test a Giesl ejector vs a Standard orifice plate DC on a BR9. The free nozzle area for the Giesl was given as 30.2 sqins, though the reported back pressures are lower than expected, more consistent with 33.2sqins free nozzle area. I am assuming, but have not proved that this is also a measurement problem. Looking at the data you quote, the backpressure for the Lemaitre is consistent with a steam rate of about 25000lbs/hr. For the Giesl to be as good as Giesl claims, the free nozzle area would have to be over 40 sqins, equivalent to a 7 ¼ “ single pipe. I do not know what 34064 had, but I’d bet it wasn’t this. So, non -comparable measurements I think. I don’t think 34064 was indicated, so Giesl’s power calculations are on the basis of what would happen if the backpressure drop was as large as claimed. The Giesl on the BR9 didn’t dramatically change perceived performance. The reason, I think is simple. The steam rates required on Newport duties for BR9s were so low as for the reduction in backpressure achieved to not be significant. Giesl couldn’t believe how little and how lightly the BR9s were used.
It’s so long since I wrote in the thread, I’m not sure which boiler programme you are referring to. I have one of my own, rough and ready but works, but it cannot answer questions of the kind you ask. I have one from Professor Hall (He was chairman of the British Heat Transfer Group, so knew what he was doing!), which he abandoned uncompleted. He was concerned about some aspects of the theory; it is also obvious that it is a heat transfer model that does not take into account all factors involved in boiler efficiency, most crucially coal loss. I had rather ignored it, but recently discovered that it makes a quite remarkable (to me at least) prediction, namely that if there were no radiant heat transfer to the superheater, superheat would be very poor indeed. Obviously there is a radiant section at the beginning of the firetubes, and it is here where the majority of superheating occurs. What is critical than is the surface area of the superheater in this radiant zone, which will be greatly influenced, as you suggest by how near the superheater gets to the firebox. Chapelon consciously got his tubes as close as he could; because of the risk of burning the elements, most others preferred to keep them some distance away to reduce maintenance, but at the cost of superheat.
On this logic, if you make the firebox bigger, you could well reduce the radiant area in the firetubes which would not be good for superheat. Looking at US designs with large combustion chambers, even with A type superheaters, the level of superheat achieved was good but not top of class, although the PRR T1 with an E type superheater was excellent. It was short between the tubeplates, and I wonder if this meant the ends of the superheater were allowed nearer the firebox. The Chapelon 4-8-0 also had a very short boiler, and I wonder if, when tubes were short, designers, who were trying to maximise superheater areas, didn’t finish up with elements nearer the firebox.
The superheat measured on the V2 at Swindon was very good, also on the BR7 boiler of similar length, even better on the even shorter BR9.
So, I think the question you put is really interesting, but Bill’s boiler model is not good enough, nor is it validated against data, to allow conclusions at the level of detail your query requires.
On your corrected comment, I’m not sure why less water in the boiler would increase evaporation? Combustion is determined by amount of coal and draught, heat transfer is independent of boiler level surely? The amount of steam/hr/ gallon water would go up with a smaller boiler, but surely not the gross amount?
Sorry there is an error in my last post.
The second sentence of the second paragraph should read: So, compared with A4 there is less water in the boiler so for a given amount of heat produced in the firebox one might expect a higher evaporation rate.
Coming to Kylchaps and A4s I have a question that your computer program might help to answer. Coster says in his book on the A4s that the development of the LNER boiler might, with the benefit of hindsight, have been developed somewhat differently. That the 41.25 sq ft was quite adequate given the Kylchap but that firebox volume might have been increased.
So have a combustion chamber 1 ft longer than in an A4 and tube length 1 ft shorter (about 17 ft) as in the V2, A1 and A2. So, compared with A4 there is less water in the boiler so for a given amount of heat produced in the firebox one might expect a higher combustion rate. Less water in boiler reduces the overall weight so boiler plates can be a little thicker and boiler pressure increased a little as Gresley planned. The possible problem is that with the rear tubeplate 1 ft further from the firegrate so also are the rear ends of the superheater tubes. And lower superheat would be very negative from the efficiency point of view, perhaps negating the point of this thought exercise. Does your program make any prediction for the superheat in such a boiler, i.e. as an A4 but with about a 2 ft long combustion chamber and a barrel as in an A1?
Thanks for info, which I was unaware of. The blastpipe nozzle of 4.5 in would give a steam rate of below the critical value of about 93 kg/(cm2 h) for 20000 lb/h, but not by much.The minimum ssc you give for this single blastpipe (15.3 lb/ihp h) seems to indicate a steam temperature of at least 330 C (626 F) for 225 lb/in2 (following a steam table in Giesel's book Locomotive Athletes for a cylinder volume of 200 L), which would need to be corrected upwards to allow for the relatively small cylinders.
Giesel also gives the results of testing the West Country class pacifics with both the Lemaitre and Giesel exhausts. In going from the Lemaitre to the Giesel, blastpipe energy fell from 165 to 78 HP lowering pressure from 0.30 to 0.13 atu, increasing pumping work from 13.1 to 15 HP, which increased smokebox vacuum from 118 to 126 mm WS, lowering shock loss from 126 to 15 HP, and increasing other losses from 25.9 to 48 HP, which was enough to lift the exhaust clear of the engine. Pumping efficiency rose from 7.8 to 19.3%.
Thanks. I think the tests you describe are very important in that they had a crucial impact of UK locomotive design ever after. I have still to locate the full report, but what ex LMS senior engineer AJ Powell reported was that (surprisingly) there was no improvement in coal or water consumption for the Kylchap set up vs an extremely restrictive 4 ½” single blastpipe. I am at a loss to explain this. He notes that the Kylchap device used was probably far too powerful for the quite small Jubilee boiler, and that there was a continuous shower of sparks from the chimney, which may explain the lack of reduction in coal.
I have done some calculations ( see below) on what the benefit of the Kylchap would be, assuming it was of the dimensions used by the LNER, which were similar to the French Compounds. There is, even at the relatively low steam rate of 20000lbs/hr, an improvement in IHP of about 8%, so you would expect an 8% reduction in water at the same power. ‘At the same power’ may be the root of the problem. Because the cylinder backpressure is lower, the steam consumption at a given cut off goes up by about 10%, so power goes up by 10%. So it may be that on these road trials, the Kylchap locomotive was producing more power, and this interfered with the economy assessment. Hopefully the report will throw light on this.
Kylchap
Standard
Steam rate, lbs/hr
Cut off, %
Blast
pipe pressure psig
IHP
SSC, lbs/ihp-hr
vacuum, ins
13700
13.6
1.0
850
16.1
3.8
14.8
3.7
820
16.7
2.7
19100
19.6
2.0
1300
14.7
5.9
21.5
7.6
1210
15.8
6.0
23100
23.7
3.0
1590
14.6
7.2
26.4
12
1420
15.3
-
26400
26.8
4.0
1780
14.9
8.0
30
16.2
1530
17.2
The LMS later tried a Kylchap on a Duchess Class Pacific, and found no benefit. Here, I think the result is explicable. The Duchesses were fitted with double plain blastpipes with a similar free nozzle area to the Kylchap; no backpressure benefit expected then. A Duchess was tested on the Rugby test plant where it was concluded that the draughting was adequate, though not brilliant. Now what better draughting such as a Kylchap does is increase the specific evaporation rate (lbs steam/sqft/hr) possible. The Duchesses had grates of 50 sqft and in service were never required to steam at more than 30000lbs/hr (600lbs/sqft/hr) because they were hand fired. On test, they got 800lbs/sqft/hr; a Kylchap can deliver 1000lbs/sqft/hr- both these higher rates are quite beyond any fireman. So because only modest specific evaporation rates were possible in service, these could be satisfactorily delivered with a very ordinary plain double blastpipe.
So, in two sets of tests, the LMS concluded that Kylchaps weren’t worth it. The LMS mafia were in charge of BR locomotive design after nationalisation and guess what- no Kylchaps. BR standard designs were hobbled with very primitive exhausts. The mafia even tried to stop the ex LNER region from fitting Kylchaps to their Pacifics, where there was a clear need.
Just shows what happens when a) you have dodgy data and b) no sound quantitative theoretical model. Politics wins.
Thanks, I'll try again. Giesel in his Anatomy of the Steam Engine (in German) mentions tests the LMS did in 1934 comparing different blastpipes on Jubilee 4-6-0s, which initially were not great steamers. In all cases steaming rate was 20000 lbs/hr. For each arrangement he gives the total power in the exhaust at the blastpipe, the pumping power (that used for draughting the boiler), the shock losses (caused when the steam impacts the flue gases in the smokebox), other losses (not totally wasted as it lifts steam out of driver's line of sight), the smokebox vacuum, pumping efficiency (ratio of pumping power to total power) and blastpipe pressure. For the standard single blastpipe the results were: 238 HP, 9.5 HP, 154 HP, 74.5 HP, 91 mm WS, 4.0% and 0.45 atu. For an improved single blastpipe: 196 HP, 11.6 HP, 126 HP, 58.4 HP, 103 mm WS, 5.9% and 0.36 atu. And for a double blastpipe: 143 HP, 13.1 HP, 94 HP, 35.9 HP, 118 mm WS, 9.1% and 0.26 atu.
One sees a progressive fall in exhaust power, improvement in draughting and also implied a fall in cylinder backpressure, giving a further gain in useful power.
Neil, if you are using Internet Explorer 9, and didn't just mistakenly forget to type a message above, you will have to look for the small grey 'compatability view' icon that looks like a torn square, meaning a sheet of paper. It is just to the right of the top URL bar, along with a magnification glass icon, the curved 'refresh view' arrow, and the close-out X. Just click on the torn sheet of paper, at which it will conform to the view intended for this server and then it will turn blue.
Welcome to the forum.
Crandell
It would be interesting ot compare these various designs differences based on actual performance. A lot of data would be needed beyond most reports of drawbar horsepower, coal & consumption, and design parameters. Things like flame, firebox, flue & tubes temperatures measured at various points; percentage of imcomplete combustion & weight of air required for combustion; stack gas temperature, weight of flow & composition analysis; steam chest pressures and superheat temperatures; continuous pressure readings throughout the cylinder cycles; etc.
Too bad a lot this testing would require a time machine...or generous funds and cooperation by someone like the UP steam program or similar group that operates a late design steam locomotive.
Given that only about 12% of the energy of the steam is converted to mechanical energy at the wheels, the focus should be to try to design a system where the total heat of the fuel fed into the firebox, the total heat produced by combustion, and the total heat absorbed by the boiler producing steam is as close to unity as possible.
"If a nation expects to be ignorant and free, it expects what never was and never will be." Thomas Jefferson
Burgard540 If you look at the chart, notice that the proportions of heating surface are nearly identical regardless of the type of locomotive. Of the total heating surface: 63% is tubes & flues, 7-8% is firebox, and 29-30% is superheater. Notice that the ALCO built UP locos had the highest proportion of firebox heating surface at 8.4% for the articulated's and 7.8% for the 4-8-4's.
If you look at the chart, notice that the proportions of heating surface are nearly identical regardless of the type of locomotive. Of the total heating surface: 63% is tubes & flues, 7-8% is firebox, and 29-30% is superheater. Notice that the ALCO built UP locos had the highest proportion of firebox heating surface at 8.4% for the articulated's and 7.8% for the 4-8-4's.
Just to add some additional data, the 2-8-8-4 EM-1, built by Baldwin in 1944-45 had its boiler comprised of 61.2% (4,540 sq.ft.) Indirect Heating Surface, 28.6% (2,118 sq.ft. Type E) Superheater Surface, and 10.2% (758 sq.ft.) Direct Heating Surface. Baldwin clearly changed the boiler design on the EM-1 from the earlier AC10 and M3/M4 designs by enlarging the size of the firebox area vs. the fire tube area.
Thanks for these additional inputs- I'm heading off for two weeks vacation now, and will add my thoughts when I'm back
GP40-2 Comparing a locomotive's total heating surface doesn't tell the full story about the steam production capacity of a boiler and must be used with caution when comparing steam locomotives.
Comparing a locomotive's total heating surface doesn't tell the full story about the steam production capacity of a boiler and must be used with caution when comparing steam locomotives.
Thanks - that's pretty much my point in plain language.
Regarding tube length, the example Ralph Johnson uses of a USRA loco where increasing the flue length from 19 to 25 ft long increased the heating surface 29% (sounds good) but only increased the heat absorption by 2%.
Again, my point is that tubes and flues are more effective heat absorbers per unit if they are of moderate length and the velocity of the combustion gases is high. However, to achieve the better convective heat absorption requires decreasing the gas flow area, thus increasing the draft and back pressure. The trade-off of slightly less engine efficieny for greater boiler efficiency. Although to me, the focus should be increasing combustion efficiency.
Burgard540 I could have written my previous post more clearly. It was more born out of the questions regarding locomotive boiler capacities and performance given their design parameters such as direct heating surface (firebox & combustion chamber), heating surface of tubes & flues, gas flow area, etc. The ALCO built steam locomotives for UP (FEF's, Challengers, and Big Boys) had significantly less heating surface compared to locomotives of similar size. However, their performance was comparable. Going just by total amount of HS doesn't explain the boiler's capacity, assuming similar heat energy inputs.
I could have written my previous post more clearly. It was more born out of the questions regarding locomotive boiler capacities and performance given their design parameters such as direct heating surface (firebox & combustion chamber), heating surface of tubes & flues, gas flow area, etc. The ALCO built steam locomotives for UP (FEF's, Challengers, and Big Boys) had significantly less heating surface compared to locomotives of similar size. However, their performance was comparable. Going just by total amount of HS doesn't explain the boiler's capacity, assuming similar heat energy inputs.
But, were they really "undersized" compared to other steam locomotives? Or, was it just a case of a different ratio of direct heating surface to indirect heating surface compared to other locomotives?
As Dreyfusshudson correctly pointed out, a locomotive with many small diameter, long tubes did not get any additional heating benefit out of the last 4 or 5 feet of those tubes. However, what the additional length of the smaller tubes did was to give the locomotive the appearance of having a larger heating surface than a locomotive with shorter, larger diameter fire tubes.
Case in point. The N&W Class A, designed in 1935-36, used 24 feet long, 3.5" diameter tubes and had an evaporative heating surface of 6,639 sq.ft. However, its firebox area (which by far produces the majority of the steam) was 587 sq.ft.
Contrast that to the B&O EM-1, which was designed in 1944. The EM-1 used 4" diameter tubes just 20 feet long for a total evaporative heating surface of 5,300 sq.ft. What that doesn't tell you is that the EM-1 had a firebox area of 758 sq.ft. which was equal to the H8 Allegheny in size. The firebox and combustion chamber on an EM-1 was 27 feet long, 7 feet longer than its fire tubes. The UP Big Boy had similar construction with a 704 sq.ft. firebox area.
Theoretically, the greater diameter of tube require less draft & less back pressure, but as the diameter increases the less heat is transferred. The length of the tube should not be beyond 100-120 internal diameters. Most of the temperature difference in the combustion gases is needed to overcome the stationary gas film along the tube's walls. The best way to reduce the films effects is to increase the gas velocity. The greater velocity of flow also causes the gases to be exposed to more tube surface in a given time, thus increasing the heat transfer of the gas. Stationary gases in the tube can only give heat via radiation which as you mentioned only occurs for a brief initial length of tubes.
So, the ideal tubes and flues should be as small and long as possible with high velocity of gas flow. Of course the there is diminishing returns, where smaller diameters and longer flues lead to greater soot build-up and increasing back pressures.
The question then becomes for a given quantity of coal burned and combustion gas produced, does a the boiler perform better with relatively equal gas flow area and higher gas velocities through the tubes of less area vs. larger amounts of tube heating surface and slower gas velocity? And at what point does increasing velocity for a given tube length lead to too high exit temperatures, greater draft & back pressure, and thus loss of efficiency? Not to forget the front end design....
Clearly, an undersized locomotive boiler could be forced and produce the same quantity of steam as a larger boiler at the sacrifice of fuel efficiency. However, I don't think steam loco designers did this on purpose, nor would a railroad like UP (which had an excellent engineering department) accept this flaw (i.e. losing money on extra needed fuel for the same train hauling capacity).
No additional comments needed regarding the lack of heat transfer the last few feet of tube length, that's been well documented through various tests.
I agree that radiative heat transfer occurs in the tubes & flues briefly. That idea had occurred to me a while ago after studying the radiation of luminous and non-luminous gases and flames. The combustion gases entering the flues and tubes would have significant amounts of CO2 and H2O (good radiators) and be at high temperatures approaching 2000 degrees F.
Regarding superheater elements too close to the firebox, the Big Boys originally had the superheater elements end 15” from the firebox tube plate. However, that was increased to 24” in March, 1943, and later increased again to 36” in August, 1950. It would be interesting to see if there were any noted performance differences or tests done to measure the effect on the final superheat temperature (or possible less pressure drop) from the shortened elements.
Sadly a lot of these questions would require full scale testing, and each locomotive would be its own special case. Needless to say, some good parameters would be found.
Cheers
Addendum - Of course not to be neglected in the design is the type of coal or fuel used.
Dear Joe,
Thanks- not sure if I understand the beginning of what you write properly. Are you saying that if you increase boiler resistance by either narrowing the tubes or making the boiler longer to improve heat transfer, you have to improve draughting, which means that you have to narrow the blastpipe, which means back pressure goes up, so in effect you are trading boiler efficiency for engine efficiency?
If so, I would say the following. Firstly, back then, they did have to make this trade off, because the understanding of front end design we have now was not available. However, Chapelon said, only half in jest I think, that you should actually start locomotive design with the front end exhaust system; with the kind of systems he was developing, you could up the draught if need be without increasing backpressure, so the trade off was not needed.
More fundamentally I would say this. According to the computer model I have, heat transfer in the last few feet of the boiler tubes is negligible; every little bit of extra evaporation/efficiency is worth having if there is no cost associated with it of course, but longer tubes do cost both in materials, maintenance, and quite probably reduced superheat- a far greater loss to engine efficiency than the very small gain in boiler efficiency the longer tubes give. So, the proper solution if faced with this would be to cut the tubes down- 18’ is more than enough I would guess. (Though it wouldn’t do much for aesthetics if the Big Boy/Challenger boilers were 5’ shorter!)
So, for this reason, I don’t think the variation in heating surface areas you quote will be of much consequence to boiler efficiency, though they may indirectly influence superheat.
I have just rediscovered some stuff that adds weight to the idea that optimising boiler tube heat transfer is not that big a target. One of our testing stations looked at their data and tried to calculate what firebox temperature ought to have been from first principles- Stefan Bolztmann and the like. The approach really only needs you to know radiant surface area (which they assumed was just the firebox), and the tube gas flow, and the amount of coal fully burned (as opposed to fired, given some is lost unburned). You get these latter two via Orsat smokebox gas analyses.
Here are some figures they quote for where the burned coal’s heat goes; firebox radiation 38.3%, firebox convection 5.5%, superheating steam 9.5%, evaporation via convection in tubes 28.1%, 18.7% lost in flue gases. Trouble is that these calculations indicate that firebox temperatures were about 300F higher than they actually measured. Now they knew enough to know that what the pyrometer said was not necessarily accurate; even so, 300F is a big difference. Since it is the T4 Stefan Bolztmann radiant contribution that has the biggest effect on the estimate of T, I turned the calculations on their head and asked ‘What is the radiant surface area if the measured temperatures are correct?’ The answer is roughly twice the actual firebox area, which implies there is a radiant section in the firetubes just less than 10% of their length- about 1.5’ on UK designs close to the firebox. Now with this kind of approach there is a huge uncertainty on what this radiant length would actually be, but I think the data are good enough to confirm that the tubes do have a radiant section, hence the radiant heat transfer is significantly greater than the 38.3% above, hence the convective tube contribution significantly less than 28.1%, and since gas temperature decreases exponentially along the tubes, no surprise that losing this small piece of convection is a minor contributor. This also supports the idea I floated earlier based on computer models, that, were it not for a radiant section of the superheater close to the firebox, superheaters would be pretty ineffective, and there is a trade off between having elements too close to the firebox and getting burned, and amount of superheat you get.
If only there was, or we could do, proper experimentation to sort all these things out!
With regard to blastpipe pressures, then at equal steam rate, assuming (not necessarily correctly) that all blastpipe discharge coefficients were the same and close to 1, and that the earlier analysis of Big Boy blastpipe diameters over time is correct, then it would be 1942 Big Boy= H-8<A< 1946 Big Boy, with the Blast going sonic (> 12.3 psig) at about 110000 lbs/hr cylinder rate on the first two, around 90000lbs/hr or less on the latter.
Some other ideas:
The gas area through the tubes and flues should be as high as possible to avoid the draft loss (especially at higher firing rates). Samller area (and longer tubes) requires greater back pressure in the cylinders, due to higher velocities in the tubes. The higher velocity has the advantage that since for convection the heat transfer is largely dependpent on the amount of gas flow, doubling the gas flow per tube increases the evaporation 90%. Total gas flow area is limited by the diameter of the boiler due to loading gauge.
So intuitively it seems that there's a trade off for greater heat transfer & evaporation due to higher gas velocities in the tubes, at the cost of higher back pressure in the cylinders. This could explain the vast disparity between the heating surface for the locomotives below.
I calculated the gas flow are for the high powered articulated loco's: Big Boys, Allegheny's, and N&W "A"s. My calculated answer for Big Boy (first series #4000-4019) was within less than 1% than the given figure of 1631 sq in or 11.33 sq ft. I assumed that the 2.25" tubes were all 11 BWG (0.120") wall thickness.
Big Boys (4000-4019) - 11.43 sq ft (Flues & Tubes 22 ft long); Total HS 8,355 sq ft
Allegheny (1600-1644) - 12.49 sq ft (Flues & Tubes 23 ft long); Total HS 10,426 sq ft
Allegheny (1645-1659) - 13.00 sq ft (Flues & Tubes 23 ft long); Total HS 9,717 sq ft
Class "A" - 11.04 sq ft (Flues & Tubes 24'1" long); Total HS 9,353 sq ft
Curious what the back pressures were for these locomotives under full power, especially the "A" since it had only 3.5" flues at over 24 ft long. Big Boys had a back pressure of 18 psi.
Going back to Big Boys lbs of water evaporated per of coal, the figures given for the April, 1943 tests were actual numbers not equivalent evaporation. The 4.44 lbs water evaporated would be increase to 6.40 lbs equivalent evaporation per lb of coal fired. Equivalent evaporation factor about 1.44. I thought of this after seeing a few old reference books rating the boilers' evaporation per pound of fuel this way.
Other thoughts?
Joe
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