1 Steam Cars : Phorum
General Steam Car topics
condensor/boiler area ratios
Posted by: tom ward (IP Logged)
Date: August 20, 2002 10:00PM
<HTML>I'm curious if there is a rule of thumb for the ratio between the heat exchange areas of condensors and boilers. I hope those who have Dobles or condensing Stanleys would provide data on those systems.
Also, what are the pros and cons of water injection condensors?</HTML>
Also, what are the pros and cons of water injection condensors?</HTML>
Re: condensor/boiler area ratios
Posted by: Jim Crank (IP Logged)
Date: August 21, 2002 02:19PM
<HTML>Tom,
As far as Dobles are concerned today, this ratio you are asking about has no relationship to anything. The average Doble Series E today has about 100 square feet of surface in the steam generator. The condenser, like all old steam cars, is not big enough to take care of everything under all conditions.
A good clean aluminum fin and tube core condenser with about 10-12 fins per inch and 4" thick will condense 100 lbs/hr per sq ft at average conditions of automotive use. So a 100 hp system with a ten pound water rate would need a 10 square foot condenser. Rough numbers; but derrived at Besler's plant on several systems and it worked.
Spraying water into the top of the condenser was all the rage back in the 1950-60's period with several SACA people with Stanleys. Worked fine and did help until the water tank got so hot the pumps started hammering.
The condenser area and the heating area all depend on the efficiency of the engine and thus the amount of steam needed per hour to provide the power one wants, and all cars are calculated individually.
Jim</HTML>
As far as Dobles are concerned today, this ratio you are asking about has no relationship to anything. The average Doble Series E today has about 100 square feet of surface in the steam generator. The condenser, like all old steam cars, is not big enough to take care of everything under all conditions.
A good clean aluminum fin and tube core condenser with about 10-12 fins per inch and 4" thick will condense 100 lbs/hr per sq ft at average conditions of automotive use. So a 100 hp system with a ten pound water rate would need a 10 square foot condenser. Rough numbers; but derrived at Besler's plant on several systems and it worked.
Spraying water into the top of the condenser was all the rage back in the 1950-60's period with several SACA people with Stanleys. Worked fine and did help until the water tank got so hot the pumps started hammering.
The condenser area and the heating area all depend on the efficiency of the engine and thus the amount of steam needed per hour to provide the power one wants, and all cars are calculated individually.
Jim</HTML>
Re: condensor/boiler area ratios
Posted by: George Nutz (IP Logged)
Date: August 21, 2002 07:17PM
<HTML>Tom,
The question should be roughly how many square feet of condensor area per gallon of fuel burned, thus leaving boiler size and engine steam rate out of the question. A 20HP Stanley of roughly 100 square feet heat transfer area may only evaporate 300-400#/hr wheras a Doble up to 2000#/hr. It is much more a function of evaporation in #/hr than anything else, hopefully one of the good persons on this great website will come up with an engine that only requires 5#/hr-HP under ideal conditions and someone else with state-of-the-art heat transfer condensors. We are all searching for the same thing!
Best, George</HTML>
The question should be roughly how many square feet of condensor area per gallon of fuel burned, thus leaving boiler size and engine steam rate out of the question. A 20HP Stanley of roughly 100 square feet heat transfer area may only evaporate 300-400#/hr wheras a Doble up to 2000#/hr. It is much more a function of evaporation in #/hr than anything else, hopefully one of the good persons on this great website will come up with an engine that only requires 5#/hr-HP under ideal conditions and someone else with state-of-the-art heat transfer condensors. We are all searching for the same thing!
Best, George</HTML>
Re: condensor/boiler area ratios
Posted by: ronald parola (IP Logged)
Date: August 21, 2002 11:45PM
<HTML>Since driving through Nevada, Utah and Colorado, we found that , yes the Stanley condenser is not big enough, Really! So what we have now is a modern IE car aluminum radiator core in place of the (mostly lead) original. Since the tanks unbolt they can be easily swapped, well easily one way the new one weighs less than a quarter of the original. It is four rows thick, I think a fifth would fit, but the the laws of dimishing returns apply, you would be cooling hot water/steam with hot air. And yes I have an electric fan behind it. With this the car doesn't leave a vapour trail except up a long hill, but you can hear the fan come on at speed, it cycles on a 195f at the outlet. Cheers Ron P</HTML>
Re: condensor/boiler area ratios
Posted by: Pat Farrell (IP Logged)
Date: August 22, 2002 03:16AM
<HTML>Ron, Did you find your old condensor plugged with steam cylinder oil? Or you were just after better efficiency? What is your best range with your 740 on a tank of water now? What was it before with the stock condensor? Just a guess would be good. Do you worry about your new condensor becoming plugged up? Do you use filters to catch your steam cylinder oil in your condinsate? I am presently running with your same original problems on my 735. Thank you.</HTML>
Re: condensor/boiler area ratios
Posted by: George Nutz (IP Logged)
Date: August 22, 2002 09:27AM
<HTML>Ron,
Your new condensor sounds great---ditto on performance info as mentioned by Pat Farrell. Could you give a make or model of the one you are using, is the original 4 rows deep?
Thanks, George</HTML>
Your new condensor sounds great---ditto on performance info as mentioned by Pat Farrell. Could you give a make or model of the one you are using, is the original 4 rows deep?
Thanks, George</HTML>
Re: condensor/boiler area ratios
Posted by: David K. Nergaard (IP Logged)
Date: August 22, 2002 10:15AM
<HTML>A suitable oil separator is not expensive. Using Stanley condensor and piston valve engine, I get 70-100 miles per tank of water. With slide valve engine it was 50-70 miles.</HTML>
Re: condensor/boiler area ratios
Posted by: Brian Drake (IP Logged)
Date: August 22, 2002 01:13PM
<HTML>Has anyone checked out using the radiators made by these guys: [www.thebrassworks.net] as a condensor? If the figures they're claiming are accurate (3.2 times the internal core surface area in direct contact with the cooling system liquid compared to a tube-and-fin radiator), then one could get away with having a smaller condensor.</HTML>
Re: condensor/boiler area ratios
Posted by: George Nutz (IP Logged)
Date: August 22, 2002 05:04PM
<HTML>Brian,
Thanks much for the info---just checked it out and wrote them on applicability to steam condensing. Last year there was a very good paper on finned tubing "EFFECT OF FINS ON BOILER PERFORMANCE" on the website
[pages.hotbot.com] but it seems no longer available.
Basically it demonstrated an increase of six fold the heat transfer per sq.ft.(a figure that Abner Doble used way back!) with about 6 fins per inch of tubing.
Thanks, George</HTML>
Thanks much for the info---just checked it out and wrote them on applicability to steam condensing. Last year there was a very good paper on finned tubing "EFFECT OF FINS ON BOILER PERFORMANCE" on the website
[pages.hotbot.com] but it seems no longer available.
Basically it demonstrated an increase of six fold the heat transfer per sq.ft.(a figure that Abner Doble used way back!) with about 6 fins per inch of tubing.
Thanks, George</HTML>
Re: condensor/boiler area ratios
Posted by: William L. Petitjean, P.E. (IP Logged)
Date: August 24, 2002 02:43PM
<HTML>Mr. Crank has hit the nail on the head - again. Condenser surface area is a function of the condenser pressure because the density of the steam at low absolute pressures governs the condenser volume and heat exchanger surface area required to develop the highest possible vacuum with a given cooling water temperature. If we go back to old Keenan and Keyes Steam Tables (Mr. Nutz's predecessors) we find that a condenser operating at 26" vacuum will contain steam that occupies 176.7 cu. ft. per lb. If, as Mr. Crank suggests, the engine consumes 10 lbs per IHP hour and produces 100 IHP then we must run about 176,700 cu. ft. per hour, or 2,945 cu. ft. per min. of steam through our condenser. This is a lot of vapor to cram through a 10 sq. ft. condenser.
However, experience proved the small condenser worked. Nonetheless, I wonder at what real absolute pressure. If the pressure in the condenser declines to, say, 21" vacuum the condenser steam will occupy 82.52 cu. ft. per lb. This is something less than half the specific volume at 26" vacuum. So, we can see that small changes in condenser vacuum cause big changes in specifc volume that materially affects the efficiency of a given heat exchange surface in a given condenser volume.
A primary drawback in small automobile condensers is their very limited volume and the resulting small heat exchange areas. This is exacerbated by the fact you are trying to transfer heat from a diffuse low specific heat vapor to an equally low specific heat gas - air. If you used the standard condenser design equations to size your automotive condenser I am quite sure you would need an RV sized trailer just to contain the properly sized condenser.
This is why I decided to use the crankcase as a condenser in my dream engine. This also means you have to abandon the traditional double acting engine and use single acting pistons with exhaust ports in the cylinder. You can still use slide valves, piston valves or poppet valves for admission, but more attention should be directed at making the cylinder see lower absolute pressures - this is where the greatest economy (water rate) improvements lurk in steam automobile power plants.
A crankcase condenser should use a jet condenser that draws the exhaust steam into the cooling water stream, condenses it and then strips the rejected heat in a remote conventional finned tube radiator. Now you are using ambient air to cool a liquid with a relatively high heat density. The small radiators required by automobiles are much more efficient in this type of service. Careful design of the jet condenser's water powered ejector will be required so improved engine efficiency is not eaten up by the ejector water pump's power requirments.
Let's talk about steam rates for a moment. I like 300 psig D&S for an inlet because that is the point on the saturation curve (see your nice Mollier Chart) where saturated steam contains the most enthalpy (total energy) per lb. This corresponds to a temperature of about 420 deg. F - a nice temperature to work with. Consultation with some more old Skinner notes shows that you can achieve a theoretical steam rate of 10.33 lbs per HP hour when expanding down to 26" vacuum (4" Hg absolute). If we go up to 600 psi at 750 deg. F the steam rate is 7.65 lbs per HP hour when expanding to 26" vacuum. This lower steam rate is 75% of the steam rate at the 300 psig saturated pressure.
However, if we operate the condenser down to 29" vacuum at the 300 psig saturated inlet pressure (29" is a common condenser pressure on steam turbines and is nearly perfect) we can realize a steam rate of 8.67 lbs per HP hour. This is 84% of the original 10.33 lb. per HP hour rate. Now the high pressure/temperature steam rate is about 88% of the improved 300 psig D&S/29" vacuum steam rate.
In my mind the scorching hot temperatures are much more difficult to deal with, both in the boilers and the engines. Furthermore, it is more difficult to insulate against 750 deg. F or higher. This means heat losses probably will eat up a lot of the 12% theoretical gain the high superheat engine should be able to attain. Finally, simple expansion engines can't utilize all that internal energy in one expansion. This means the exhaust will be superheated in proportion to the inlet temperatures at a given cutoff and the condenser will overload even more.
In conclusion, special attention to condenser design not only will help achieve 100% condensate return to the boiler, it will materially improve the steam rate of most any well designed steam prime mover - a double benefit. Oh, by the way if my hot head engine could be made to achieve 8.65 lbs per IHP hour, put out 20 IHP and exhaust to a nice crankcase condenser with low exhaust loss my injection water volume per stroke will decline to about 0.011 cu. in. per injection "spurt". Oops! I think I made an error on my previous posts. At the Skinner water rate of about 11.4 lbs. per IHP hour the injector "spurt" is about 0.015 cu. in., not 0.9 cu. in. This is much more doable. But, as I said above, the key is proper condenser/engine design synergy to get that last bit of internal energy out of the steam before throwing away the latent heat in the condenser.
This is not exactly a rule of thumb, but it was fun writing it anyway.
Bill Petitjean</HTML>
However, experience proved the small condenser worked. Nonetheless, I wonder at what real absolute pressure. If the pressure in the condenser declines to, say, 21" vacuum the condenser steam will occupy 82.52 cu. ft. per lb. This is something less than half the specific volume at 26" vacuum. So, we can see that small changes in condenser vacuum cause big changes in specifc volume that materially affects the efficiency of a given heat exchange surface in a given condenser volume.
A primary drawback in small automobile condensers is their very limited volume and the resulting small heat exchange areas. This is exacerbated by the fact you are trying to transfer heat from a diffuse low specific heat vapor to an equally low specific heat gas - air. If you used the standard condenser design equations to size your automotive condenser I am quite sure you would need an RV sized trailer just to contain the properly sized condenser.
This is why I decided to use the crankcase as a condenser in my dream engine. This also means you have to abandon the traditional double acting engine and use single acting pistons with exhaust ports in the cylinder. You can still use slide valves, piston valves or poppet valves for admission, but more attention should be directed at making the cylinder see lower absolute pressures - this is where the greatest economy (water rate) improvements lurk in steam automobile power plants.
A crankcase condenser should use a jet condenser that draws the exhaust steam into the cooling water stream, condenses it and then strips the rejected heat in a remote conventional finned tube radiator. Now you are using ambient air to cool a liquid with a relatively high heat density. The small radiators required by automobiles are much more efficient in this type of service. Careful design of the jet condenser's water powered ejector will be required so improved engine efficiency is not eaten up by the ejector water pump's power requirments.
Let's talk about steam rates for a moment. I like 300 psig D&S for an inlet because that is the point on the saturation curve (see your nice Mollier Chart) where saturated steam contains the most enthalpy (total energy) per lb. This corresponds to a temperature of about 420 deg. F - a nice temperature to work with. Consultation with some more old Skinner notes shows that you can achieve a theoretical steam rate of 10.33 lbs per HP hour when expanding down to 26" vacuum (4" Hg absolute). If we go up to 600 psi at 750 deg. F the steam rate is 7.65 lbs per HP hour when expanding to 26" vacuum. This lower steam rate is 75% of the steam rate at the 300 psig saturated pressure.
However, if we operate the condenser down to 29" vacuum at the 300 psig saturated inlet pressure (29" is a common condenser pressure on steam turbines and is nearly perfect) we can realize a steam rate of 8.67 lbs per HP hour. This is 84% of the original 10.33 lb. per HP hour rate. Now the high pressure/temperature steam rate is about 88% of the improved 300 psig D&S/29" vacuum steam rate.
In my mind the scorching hot temperatures are much more difficult to deal with, both in the boilers and the engines. Furthermore, it is more difficult to insulate against 750 deg. F or higher. This means heat losses probably will eat up a lot of the 12% theoretical gain the high superheat engine should be able to attain. Finally, simple expansion engines can't utilize all that internal energy in one expansion. This means the exhaust will be superheated in proportion to the inlet temperatures at a given cutoff and the condenser will overload even more.
In conclusion, special attention to condenser design not only will help achieve 100% condensate return to the boiler, it will materially improve the steam rate of most any well designed steam prime mover - a double benefit. Oh, by the way if my hot head engine could be made to achieve 8.65 lbs per IHP hour, put out 20 IHP and exhaust to a nice crankcase condenser with low exhaust loss my injection water volume per stroke will decline to about 0.011 cu. in. per injection "spurt". Oops! I think I made an error on my previous posts. At the Skinner water rate of about 11.4 lbs. per IHP hour the injector "spurt" is about 0.015 cu. in., not 0.9 cu. in. This is much more doable. But, as I said above, the key is proper condenser/engine design synergy to get that last bit of internal energy out of the steam before throwing away the latent heat in the condenser.
This is not exactly a rule of thumb, but it was fun writing it anyway.
Bill Petitjean</HTML>
Re: condensor/boiler area ratios
Posted by: George Nutz (IP Logged)
Date: August 24, 2002 05:37PM
<HTML>Bill,
One of the great disadvantages of starting the expansion process at rather low temperatures and pressures(300psi) is that to achieve a low steam rate one HAS to have high condensor vacuum in order not to have backpressure on a long expansion engine. By the way the highest enthalpy point for saturated steam is around 425-450psia. From that starting point it does not require that many expansions before the end-expansion pressure is higher than the actual back pressure. In the case you mentioned in the range of 26-29", ( 2-1psia) vacuum(only achievable in boats and plants with cooling ponds) one would have to have condensing temperatures as low as 100-150 degreesF, hardly achievable in quantity on a hot summer road---the cooling air entering could be that high or higher. As the heat transfer per square foot of condensor area would go down with the lessening temperature differential a much larger condensor would be required. A case in point is that the "F" Doble engine would only require 10-12# steam/BHP-hr(notIHP) but would jump to 16-18#/BHP-hr with the same steam temperature if the pressure dropped towards 200psig---the end of the expansion cycle in the low pressure cylinder would not be doing any useful work. The Doble could achieve this 10-12#/BHP-hr even with 25psia backpressure from the exhaust turbine booster and condensor. The steam rate of a Doble and even a Stanley starts moving up quickly even if the steam inlet temperature drops by 100 degrees F. A saturated steam engine of modest steam rate is great and obtainable if one has sufficient cooling available for such very high vacuums , this may not be the case in an automobile. The condensor does not worry as much/if condensors worry at all/ about the amount of volume(and velocity) going thru it but on how many pounds of steam and BTU's it has to remove and again this becomes very difficult when vacuum condensing temperatures becomes close to the ambient air coolant temperature. Of course in your single acting engine example the crankcase would have to have this vacuum applied to it to get useful work out of the expansion cycle and not substantially raise the water rate to unacceptable levels.
George</HTML>
One of the great disadvantages of starting the expansion process at rather low temperatures and pressures(300psi) is that to achieve a low steam rate one HAS to have high condensor vacuum in order not to have backpressure on a long expansion engine. By the way the highest enthalpy point for saturated steam is around 425-450psia. From that starting point it does not require that many expansions before the end-expansion pressure is higher than the actual back pressure. In the case you mentioned in the range of 26-29", ( 2-1psia) vacuum(only achievable in boats and plants with cooling ponds) one would have to have condensing temperatures as low as 100-150 degreesF, hardly achievable in quantity on a hot summer road---the cooling air entering could be that high or higher. As the heat transfer per square foot of condensor area would go down with the lessening temperature differential a much larger condensor would be required. A case in point is that the "F" Doble engine would only require 10-12# steam/BHP-hr(notIHP) but would jump to 16-18#/BHP-hr with the same steam temperature if the pressure dropped towards 200psig---the end of the expansion cycle in the low pressure cylinder would not be doing any useful work. The Doble could achieve this 10-12#/BHP-hr even with 25psia backpressure from the exhaust turbine booster and condensor. The steam rate of a Doble and even a Stanley starts moving up quickly even if the steam inlet temperature drops by 100 degrees F. A saturated steam engine of modest steam rate is great and obtainable if one has sufficient cooling available for such very high vacuums , this may not be the case in an automobile. The condensor does not worry as much/if condensors worry at all/ about the amount of volume(and velocity) going thru it but on how many pounds of steam and BTU's it has to remove and again this becomes very difficult when vacuum condensing temperatures becomes close to the ambient air coolant temperature. Of course in your single acting engine example the crankcase would have to have this vacuum applied to it to get useful work out of the expansion cycle and not substantially raise the water rate to unacceptable levels.
George</HTML>
Re: condensor/boiler area ratios
Posted by: ronald parola (IP Logged)
Date: August 25, 2002 04:58PM
<HTML>No the condenser was very clean inside, I guess we run a high speeds long enough to keep it ALL hot and cleaned out. The BEST we have ever gotten was approx 130 miles on one tank, this was straight freeway driving, NO hills, but at high speed; well 40-45. Unfortunately we have not had such a long flat run since, but driving down to Santa Barbara on Hwy 1 which is slower but hilly we got the same water mileage, which is an improvment. Also driving to the steamboat meet from Geyserville (hilly) to Isleton (flat) we used to have to stop for water at 70 miles, now we can safely make it there. I do have a oil separator in the return (steam) line under the passenger side rocker panel, and have filter medium in the front 3rd of the water tank, yes some oil STILL gets back to the boiler but nowhere near the amounts that there was. I open the separator cocks and they drip oil for DAYS, I also drain the water tank when hot and you can see the sheen from the oil, BUT at least now I don't have the 1 inch build up of oil at the bottom of the water tank any more. And as a side benefit with the alloy condenser the steering got lighter! Cheers Ron P</HTML>
Re: condensor/boiler area ratios
Posted by: ronald parola (IP Logged)
Date: August 25, 2002 05:05PM
<HTML>The original is four rows deep. There are a number of radiator cores that are the correct size but none that thick. Since I was sure that I couldn't weld two together (I CAN'T weld aluminum well at all, hmmm a new tig perhaps?) I had one built by a race shop, Ron Dennis I believe, I can look it up if you wish. But if you can weld, I think a Ford Taurus V6 rad was the correct size, I can find out since the rad shop we deal with has size charts. Ron P</HTML>
Re: condensor/boiler area ratios
Posted by: ronald parola (IP Logged)
Date: August 25, 2002 05:07PM
<HTML>Yep, but they look too different for an old Stanley. Ron P</HTML>
Re: condensor design
Posted by: Peter Heid (IP Logged)
Date: August 25, 2002 11:32PM
<HTML>In my books on heat exchanger design, I have noticed that most automotive radiators are designed for a 15 to 20 degree air temperature differential. It would seem a higher differential might be better because of steams higher temperature and so that preheating of the air to the burner could take advantage of some of the extra heat evolving from the condensor. I believe preheating of intake air is often overlooked in modern steam designs and to recycle the heat back into the fire is one of the best things you can do with it. A vee front radiator such as that developed for use with Sterling engines has a claimed efficiency 4 times that of a straight front type.
Peter Heid</HTML>
Peter Heid</HTML>
Re: condensor design
Posted by: tom ward (IP Logged)
Date: August 27, 2002 08:40PM
<HTML>Venting airflow from the condensor to the burner/boiler has been my plan all along. This would require a fan of some sort which must be running before the burner ignites. I wonder why Doble didn't do it?</HTML>
Re: condensor design
Posted by: Terry Williams (IP Logged)
Date: August 27, 2002 10:23PM
<HTML>Can your combustion chamber materials stand up to the increased combustion temperature?</HTML>
Re: condensor design
Posted by: Peter Brow (IP Logged)
Date: August 28, 2002 04:32AM
<HTML>Hi Tom,
Actually all the condensing steam cars of the classic era, to my knowledge, fed hot air from the condenser into the burner. Their burners simply drew in radiator-heated air under the hood. I'm not sure a duct is needed, unless the boiler is in the rear of the car with the condenser in the front, or vice versa.
Peter</HTML>
Actually all the condensing steam cars of the classic era, to my knowledge, fed hot air from the condenser into the burner. Their burners simply drew in radiator-heated air under the hood. I'm not sure a duct is needed, unless the boiler is in the rear of the car with the condenser in the front, or vice versa.
Peter</HTML>
Re: condensor design
Posted by: Jim Crank (IP Logged)
Date: August 28, 2002 01:20PM
<HTML>Peter H.,
Air preheating is definitely an aid to higher efficiency.
The French-Coats design used the edges of the condenser fan as the air blower to the burner, and a couple of Dobles had air jackets around the firebox. Lears also used small aluminum condensers as air preheaters in one of their numerous flayings around, with the turbine system.
A small condenser in the air ducting fed directly from the engine, and in series with the main condenser, would be a help.
Tom,
Doble did do this in one car, D-1, where the entire auxiliary unit and burner was mounted in front of the boiler. Then they rearranged it and did not follow up on the idea of air preheat to the burner.
Terry,
Sure if you don't try the cheap approach and use the right metal for the combustion chamber.
Jim</HTML>
Air preheating is definitely an aid to higher efficiency.
The French-Coats design used the edges of the condenser fan as the air blower to the burner, and a couple of Dobles had air jackets around the firebox. Lears also used small aluminum condensers as air preheaters in one of their numerous flayings around, with the turbine system.
A small condenser in the air ducting fed directly from the engine, and in series with the main condenser, would be a help.
Tom,
Doble did do this in one car, D-1, where the entire auxiliary unit and burner was mounted in front of the boiler. Then they rearranged it and did not follow up on the idea of air preheat to the burner.
Terry,
Sure if you don't try the cheap approach and use the right metal for the combustion chamber.
Jim</HTML>
Re: condensor design
Posted by: Terry Williams (IP Logged)
Date: August 28, 2002 02:21PM
<HTML>Yer right, Jim. I've done that several times. I have a section of Inconel in my outboard just below the waterwall that is holding up fairly well, it should last a couple thousand hours under current conditions. I have been thinking of using about 400 degrees F. of air preheat in the next version which may call for a little more thought in that area.</HTML>
Re: condensor design
Posted by: Pat Farrell (IP Logged)
Date: August 28, 2002 02:22PM
<HTML>With our 1914 Stanley 606, on a hot day of about 80~F you can easily cruise at 65 M.P.H. On a cold day of about 32~F, you would be lucky to hit 50 M.P.H The hot air definitely aids the performance of our Stanley. Cold air into the mixing tubes goes right to the heart of the boiler and puts a chill to the steam. Preheated air would make a big difference.</HTML>
Re: condensor/boiler area ratios
Posted by: tom ward (IP Logged)
Date: August 30, 2002 08:11AM
<HTML>I plan to use a turbine combustor. I'm also going to experiment with carbon fiber/ ceramic boiler parts.</HTML>
Re: condensor/boiler area ratios
Posted by: William L. Petitjean, P.E. (IP Logged)
Date: September 03, 2002 05:17AM
<HTML>Hi George:
You have made some very good points. It looks like this old stationary/locomotive engine guy has a few things to learn about steam automobiles! First off, I should take my own advice and actually look at my Mollier Chart. You are correct about the peak of the saturation curve at about 425 psia. I think I got stuck on 300 psig because that is about where the locomotive designers figured they maximized their efficiencies with fairly long cutoffs in single expansion locomotives.
In regards to the maintenance of high vacuums you are also correct that the weather conditions have to be cool enough to get the very low absolute exhaust pressures in a properly designed condenser. If I recall correctly 29" of vacuum will maintain 100% steam vapor at something like 75 deg. F.
One of the reasons I favor a crankcase condenser over the traditional vapor/air arrangement is that I can condense the steam in a multi-jet condenser, create high vacuum in the crankcase and then reject the waste heat in a radiator much the same as in an infernal combustion engine cooling system -- liquid water at relatively high volumes and velocities in a forced air radiator. There is lots of experience with high efficiency liquid radiators.
South African Railways ran some condensing locomotives across their desert sections. The condensers were huge tender mounted units that directly condensed the steam with big air blowers. They were not very successful because the heat transfer efficiency was rather dismal.
I looked up a Koerting Multi Jet Condenser in my old Steam Engineering book and found out one test result showed 1.87" Hg. absolute pressure was maintained with 76 deg. F injection water corresponding to a hotwell temperature of 88 deg. F. Unfortunately, I also found out it took 87.5 lbs of injection water at 12 psi for every 1 lb of steam condensed. This means in a 20 hp engine with 10 lbs/HP hr. steam rate you would have to circulate about 34 gallons per minute which seems like a fairly high rate. However, I think a modern automobile IC power plant might be in this same range because they run fairly low delta temp. in their radiators. The one advantage with the multi jet condenser is that you don't need a separate vacuum pump because the ejector entrains all the non-condensibles.
In any case an air cooled liquid radiator can operate efficiently at much lower delta temp. than an air cooled vapor radiator. When the weather gets too hot to maintain the higher vacuums the engine needs to adapt with increasing excess clearance to accomodate the higher backpressures without sacrificing too much heat rate.
One thing that the steam tables show to advantage for the higher pressures and temperatures is the significantly lower latent heat of vaporization at higher pressures. For example, the latent heat at 300 psia D&S is 809 btu/lb. At 800 psia D&S the latent heat is 689 btu/lb. This means less load on the condenser and boiler for a given steam rate and is an important avenue to higher economy.
Nonetheless, I wonder how good the average steam rate really is in a steam car over the whole load range seen on the open road. I assume that the engine must be sized for a high horsepower output to achieve acceptable acceleration, but only a fraction of this horsepower is needed at steady state speeds even as high as 60 or 70 mph. If this is the case then the engine output must be modulated either by precise cutoff control or the inlet must be throttled. If the inlet is throttled then you are spending most of your road time at the 200 psi you mentioned or something like that and your real steam rate is closer to the 16-18 lb/BHP hr. you mentioned. The ultra high boiler pressure does you no good unless you have the thing wide open all the time (which corresponds to how most yuppies drive in the Seattle area).
Going back to my steam rate tables (which I just noticed are in lbs/kW hr. -- careful with the units!!) the best theoretical steam rates are: 8.23 lbs/IHP hr. for 1200 psi, 800 deg. F and 25 psi exhaust; 8.84 lbs/IHP hr. for 800 psi, 800 deg. F and 25 psi exhaust; 12.75 lbs/IHP hr. for 200 psi, 800 deg. F and 25 psi exhaust. Clearly, throttling the inlet to modulate engine output places a major penalty on power plant efficiency.
In the Doble car the compound engine carries another penalty because its efficiency only peaks when the HP and LP indicator cards piggyback to represent one continuous card with a single, unbroken expansion line. This is fixed by the cylinder volume ratios and is tuned by the relative cutoffs in each stage. Compounds exhibit very nice efficiencies at long cutoffs when operating near their design power range but completely go to pot when outputs are modulated outside a very narrow range.
In marine practice, Skinner Unaflows operating on single expansion arrangements with short cutoffs easily beat the water rates of triple expansion engines all the time. Because they could go to 90% cutoff they also managed about 300% overloads as well.
Now that I have my units straight I can go back to the lower pressure saturated engine. The best theoretical steam rate is 11.5 lbs/IHP hr. for 300 psi D&S and 0 psi atmospheric exhaust; 7.71 lbs/IHP hr. for 300 psi D&S and 26" Hg vacuum (4" Hg absolute); 6.47 lbs/IHP for 300 psi D&S and 29" Hg vacuum (1" Hg absolute). The best rate above is on the order of half the Doble's throttled rate at 25 psi exhaust. The worst rate at atmospheric exhaust when the condensate is about 200 deg. F is still better than the Doble's high backpressure rate at the throttled condition. Furthermore, our whole steam circuit never sees any temperature over 418 deg. F. This means we have huge delta temps to drive BTU's across heating surfaces with corresponding reduction in size. Materials and lubrication become child's play compared to the needs of the ultra high temperature monsters (namely we can use nonlube plastics everywhere).
The key, of course is to come up with an engine that effectively uses the steam's low pressure energy and is not throttled, but is controlled strictly with precise cutoff control. If you have ever been around a Skinner generator drive where the engine is modulated only with the cutoff governor you know what I mean. They are a beautiful sight and very simple to operate. The Skinner steam rate curves are long and very flat. They only turn up on the lower outputs when friction and windage become a large part of the total output.
That is why I champion some type of direct connected steam prime mover that operates on the diesel principle of injection, cutoff and expansion. I have only the utmost respect for the engineers at Skinner and Doble because they pushed the traditional steam engine to its ultimate limits. But, in today's world those limits are not far enough. Somebody has to do something magical to achieve steam rates close to the theoretical rates posted above in everyday driving or we are all just practicing a nice hobby that has no bigger application.
Finally, I really thank you George for forcing me to sharpen my pencil. I have been practicing engineering for a long time with a crayon and it feels good to get my ears boxed by a competent steam power engineer. I hope my rash, iconoclastic behavior does not upset the forum too much.
Bill Petitjean</HTML>
You have made some very good points. It looks like this old stationary/locomotive engine guy has a few things to learn about steam automobiles! First off, I should take my own advice and actually look at my Mollier Chart. You are correct about the peak of the saturation curve at about 425 psia. I think I got stuck on 300 psig because that is about where the locomotive designers figured they maximized their efficiencies with fairly long cutoffs in single expansion locomotives.
In regards to the maintenance of high vacuums you are also correct that the weather conditions have to be cool enough to get the very low absolute exhaust pressures in a properly designed condenser. If I recall correctly 29" of vacuum will maintain 100% steam vapor at something like 75 deg. F.
One of the reasons I favor a crankcase condenser over the traditional vapor/air arrangement is that I can condense the steam in a multi-jet condenser, create high vacuum in the crankcase and then reject the waste heat in a radiator much the same as in an infernal combustion engine cooling system -- liquid water at relatively high volumes and velocities in a forced air radiator. There is lots of experience with high efficiency liquid radiators.
South African Railways ran some condensing locomotives across their desert sections. The condensers were huge tender mounted units that directly condensed the steam with big air blowers. They were not very successful because the heat transfer efficiency was rather dismal.
I looked up a Koerting Multi Jet Condenser in my old Steam Engineering book and found out one test result showed 1.87" Hg. absolute pressure was maintained with 76 deg. F injection water corresponding to a hotwell temperature of 88 deg. F. Unfortunately, I also found out it took 87.5 lbs of injection water at 12 psi for every 1 lb of steam condensed. This means in a 20 hp engine with 10 lbs/HP hr. steam rate you would have to circulate about 34 gallons per minute which seems like a fairly high rate. However, I think a modern automobile IC power plant might be in this same range because they run fairly low delta temp. in their radiators. The one advantage with the multi jet condenser is that you don't need a separate vacuum pump because the ejector entrains all the non-condensibles.
In any case an air cooled liquid radiator can operate efficiently at much lower delta temp. than an air cooled vapor radiator. When the weather gets too hot to maintain the higher vacuums the engine needs to adapt with increasing excess clearance to accomodate the higher backpressures without sacrificing too much heat rate.
One thing that the steam tables show to advantage for the higher pressures and temperatures is the significantly lower latent heat of vaporization at higher pressures. For example, the latent heat at 300 psia D&S is 809 btu/lb. At 800 psia D&S the latent heat is 689 btu/lb. This means less load on the condenser and boiler for a given steam rate and is an important avenue to higher economy.
Nonetheless, I wonder how good the average steam rate really is in a steam car over the whole load range seen on the open road. I assume that the engine must be sized for a high horsepower output to achieve acceptable acceleration, but only a fraction of this horsepower is needed at steady state speeds even as high as 60 or 70 mph. If this is the case then the engine output must be modulated either by precise cutoff control or the inlet must be throttled. If the inlet is throttled then you are spending most of your road time at the 200 psi you mentioned or something like that and your real steam rate is closer to the 16-18 lb/BHP hr. you mentioned. The ultra high boiler pressure does you no good unless you have the thing wide open all the time (which corresponds to how most yuppies drive in the Seattle area).
Going back to my steam rate tables (which I just noticed are in lbs/kW hr. -- careful with the units!!) the best theoretical steam rates are: 8.23 lbs/IHP hr. for 1200 psi, 800 deg. F and 25 psi exhaust; 8.84 lbs/IHP hr. for 800 psi, 800 deg. F and 25 psi exhaust; 12.75 lbs/IHP hr. for 200 psi, 800 deg. F and 25 psi exhaust. Clearly, throttling the inlet to modulate engine output places a major penalty on power plant efficiency.
In the Doble car the compound engine carries another penalty because its efficiency only peaks when the HP and LP indicator cards piggyback to represent one continuous card with a single, unbroken expansion line. This is fixed by the cylinder volume ratios and is tuned by the relative cutoffs in each stage. Compounds exhibit very nice efficiencies at long cutoffs when operating near their design power range but completely go to pot when outputs are modulated outside a very narrow range.
In marine practice, Skinner Unaflows operating on single expansion arrangements with short cutoffs easily beat the water rates of triple expansion engines all the time. Because they could go to 90% cutoff they also managed about 300% overloads as well.
Now that I have my units straight I can go back to the lower pressure saturated engine. The best theoretical steam rate is 11.5 lbs/IHP hr. for 300 psi D&S and 0 psi atmospheric exhaust; 7.71 lbs/IHP hr. for 300 psi D&S and 26" Hg vacuum (4" Hg absolute); 6.47 lbs/IHP for 300 psi D&S and 29" Hg vacuum (1" Hg absolute). The best rate above is on the order of half the Doble's throttled rate at 25 psi exhaust. The worst rate at atmospheric exhaust when the condensate is about 200 deg. F is still better than the Doble's high backpressure rate at the throttled condition. Furthermore, our whole steam circuit never sees any temperature over 418 deg. F. This means we have huge delta temps to drive BTU's across heating surfaces with corresponding reduction in size. Materials and lubrication become child's play compared to the needs of the ultra high temperature monsters (namely we can use nonlube plastics everywhere).
The key, of course is to come up with an engine that effectively uses the steam's low pressure energy and is not throttled, but is controlled strictly with precise cutoff control. If you have ever been around a Skinner generator drive where the engine is modulated only with the cutoff governor you know what I mean. They are a beautiful sight and very simple to operate. The Skinner steam rate curves are long and very flat. They only turn up on the lower outputs when friction and windage become a large part of the total output.
That is why I champion some type of direct connected steam prime mover that operates on the diesel principle of injection, cutoff and expansion. I have only the utmost respect for the engineers at Skinner and Doble because they pushed the traditional steam engine to its ultimate limits. But, in today's world those limits are not far enough. Somebody has to do something magical to achieve steam rates close to the theoretical rates posted above in everyday driving or we are all just practicing a nice hobby that has no bigger application.
Finally, I really thank you George for forcing me to sharpen my pencil. I have been practicing engineering for a long time with a crayon and it feels good to get my ears boxed by a competent steam power engineer. I hope my rash, iconoclastic behavior does not upset the forum too much.
Bill Petitjean</HTML>
Re: condensor/boiler area ratios
Posted by: Jim Crank (IP Logged)
Date: September 03, 2002 02:30PM
<HTML>Bill,
Now you are thinking on the right track.
A compound is only effective if you are at the exact speed where the two cards mesh at that specific cutoff percentage. This goes for the Doble and the White. Otherwise it is one big compromise and the water rate suffers.
Only one car that I know about had a valve gear that independently adjusted both the HP and the LP as the speed went up, via the cutoff control. The British Lamplough-Albany of 1903. Absolutely fiendish collection of levers and pivot points; but the idea was sound.
George Nutz and I have batted this one around a lot, and our conclusion is that in a car, you are really better off with a simple three or four cylinder poppet valve unaflow with short cutoff at high speed.
Let alone Doble's triple for Leslie Hills E-12 and the proposed quadruple expansion engine Dr. Mudd sponsored in 1928. The triple was good at average cruising speed; but hard to start and thank heavens the quad was never built.
All this makes the three cylinder six piston opposed unaflow design, with poppet inlets a most attractive engine for a car. Something I was pasionate about twenty years ago.
Jim</HTML>
Now you are thinking on the right track.
A compound is only effective if you are at the exact speed where the two cards mesh at that specific cutoff percentage. This goes for the Doble and the White. Otherwise it is one big compromise and the water rate suffers.
Only one car that I know about had a valve gear that independently adjusted both the HP and the LP as the speed went up, via the cutoff control. The British Lamplough-Albany of 1903. Absolutely fiendish collection of levers and pivot points; but the idea was sound.
George Nutz and I have batted this one around a lot, and our conclusion is that in a car, you are really better off with a simple three or four cylinder poppet valve unaflow with short cutoff at high speed.
Let alone Doble's triple for Leslie Hills E-12 and the proposed quadruple expansion engine Dr. Mudd sponsored in 1928. The triple was good at average cruising speed; but hard to start and thank heavens the quad was never built.
All this makes the three cylinder six piston opposed unaflow design, with poppet inlets a most attractive engine for a car. Something I was pasionate about twenty years ago.
Jim</HTML>
Re: condensor/boiler area ratios
Posted by: George Nutz (IP Logged)
Date: September 03, 2002 04:26PM
<HTML>Bill,
Very sorry if my posts imply that I do not use crayons myself, I would never want to "box" someone in any fashion, have made a huge amount of mistakes over the years but that is the way old engineers learn I guess. I certainly welcome your considerable insights as your knowledge covers many more steam related areas than mine--please don't ever take me too seriously!!
I agree that the Skinner unaflow was a magnificient engine, much prefer the very simple unaflow for all purpose running than a triple with reheat between each cylinder. The steam rates as set by the Williams are to be respected and lower than any of Abners compound engines. By the way talked to Hal Fuller yesterday and he is still hanging on at the old Skinner facility part time, a truly creative engineer if there ever was one. What a shame of its demise--a once great and proud company.
As Jim mentioned a three cylinder DA modified unaflow would be an excellent engine, very smooth torque curves even at smaller admissions and great overload capacity(as you mentioned about the Skinner engines), especially if one incorporates the Williams automatic compression relief valve and use very small clearance volumes. Isn't this website a breathe of fresh steam for all of us!!
Well I have to go back to my crayons and make some crude drawings of a 2 cylinder "V" engine that demands attention. Please wish me luck.
Best, George</HTML>
Very sorry if my posts imply that I do not use crayons myself, I would never want to "box" someone in any fashion, have made a huge amount of mistakes over the years but that is the way old engineers learn I guess. I certainly welcome your considerable insights as your knowledge covers many more steam related areas than mine--please don't ever take me too seriously!!
I agree that the Skinner unaflow was a magnificient engine, much prefer the very simple unaflow for all purpose running than a triple with reheat between each cylinder. The steam rates as set by the Williams are to be respected and lower than any of Abners compound engines. By the way talked to Hal Fuller yesterday and he is still hanging on at the old Skinner facility part time, a truly creative engineer if there ever was one. What a shame of its demise--a once great and proud company.
As Jim mentioned a three cylinder DA modified unaflow would be an excellent engine, very smooth torque curves even at smaller admissions and great overload capacity(as you mentioned about the Skinner engines), especially if one incorporates the Williams automatic compression relief valve and use very small clearance volumes. Isn't this website a breathe of fresh steam for all of us!!
Well I have to go back to my crayons and make some crude drawings of a 2 cylinder "V" engine that demands attention. Please wish me luck.
Best, George</HTML>
Re: condensor/boiler area ratios
Posted by: William L. Petitjean, P.E. (IP Logged)
Date: September 04, 2002 01:31AM
<HTML>Jim:
I am quite excited about the three cylinder, six piston uniflow since I looked at the Comers/Rootes engine picture. I think you are exactly right about the arrangement. Even though I still like my wild hare hot head/injector idea the single poppet admission valve would make a very nice inlet system. I have done some work on solenoid valve actuation and believe an electronic valve gear would work a small, balanced valve quite nicely with very crisp cutoff at the high expansion ratios needed to get good efficiency. It would be quite simple and relatively cheap to monitor crank angle and integrate the cutoffs required for a given output.
My electronics designer went over the basics with me on a gas engine project I would like to build in the near future. PC boards may not be as nice as elegant mechanical systems, but boy are they easy to troubleshoot! Just throw the old one away and put a new one in place. Don't like the valve events? Just reprogram the processer and you have all new valve events!
The large exhaust port areas in an OP design coupled with the spacious crankcase would make a very nice crankcase condenser and water lubrication of Orkot bearings on the running gear would eliminate the old water/oil problems of long ago.
I think this type of engine could really be a winner.
Bill Petitjean</HTML>
I am quite excited about the three cylinder, six piston uniflow since I looked at the Comers/Rootes engine picture. I think you are exactly right about the arrangement. Even though I still like my wild hare hot head/injector idea the single poppet admission valve would make a very nice inlet system. I have done some work on solenoid valve actuation and believe an electronic valve gear would work a small, balanced valve quite nicely with very crisp cutoff at the high expansion ratios needed to get good efficiency. It would be quite simple and relatively cheap to monitor crank angle and integrate the cutoffs required for a given output.
My electronics designer went over the basics with me on a gas engine project I would like to build in the near future. PC boards may not be as nice as elegant mechanical systems, but boy are they easy to troubleshoot! Just throw the old one away and put a new one in place. Don't like the valve events? Just reprogram the processer and you have all new valve events!
The large exhaust port areas in an OP design coupled with the spacious crankcase would make a very nice crankcase condenser and water lubrication of Orkot bearings on the running gear would eliminate the old water/oil problems of long ago.
I think this type of engine could really be a winner.
Bill Petitjean</HTML>
Re: condensor/boiler area ratios
Posted by: Martin Werbeck (IP Logged)
Date: September 13, 2002 02:00PM
<HTML>Hello Tom,
I have discovered the excellent SACA page only a couple of weeks ago, although
I have been intersted in steam and steam cars since several years.
Thanks to your reply on the IAV- engine.
I found some figures about condensing steam locos, which might be helpful.
There were 207 German 2-10-0 condensing locos , which were build from 1943 to 1949 and used until 1954. They had boilers for 10t/hr steam rate for 1500 ihp.
The exhaust turbine driving the fans had 200 horsepower, steam coming in at 1.6 atm and out at 1.02 atm.
The tenders had ten or twelve condensing units of 230 m² (2475 ft²) cooling surface each. The engines ran sucsessfully at 46 °C air temp. and gave ranges of
at least 1000 km (600 miles).
The much larger 4-8-4 SAR engines had boilers for around 3000 ihp.
The exhaust steam drove the draft turbine of 150 hp, then the fan turbine of 550 hp,
the back pressure being about the same as in a conventional loco (1 atm=15psi).
The five 7 ft- fans ran at 1000 rpm and blew 360 kg (800lb) per second , or
13,000,000 ft³ per hour of air through the fins to condense 24 tons/hr steam at
40 °C air temp.
One of the problems was the amount of lubricating oil in the feedwater. It was considered that 5 ppm or less was not dangerous. Sometimes peak amounts of
50 ppm were measured, although the 2-10-0s showed never more than 1-3 ppm.
After some experimenting it was found that the boilers tend to prime and the water mixed with the cylinder oil forming an emulsion. But the oil separators could only work if the oil was still a liquid and the steam a gas, the heavier oil drops were held back by a series of sheets, the steam streaming around them.
The solution was a "Vortex" water separator placed between the cylinders and
the oil separator. This let the oily water dripple to the ground, and the amount
of lub oil was never higher than 2-3 ppm thereafter.
Does anyone know how much oil a Stanley should have in the exhaust steam?
According to the figures I have, it should be 1000 ppm, that can`t be, can it ??
I know that the monotube boilers from the Doble-Henschel trucks and buses
were regularly sandblasted internally to clean them from carbon deposits.
Was or is this a usual practice in America ?
Martin</HTML>
I have discovered the excellent SACA page only a couple of weeks ago, although
I have been intersted in steam and steam cars since several years.
Thanks to your reply on the IAV- engine.
I found some figures about condensing steam locos, which might be helpful.
There were 207 German 2-10-0 condensing locos , which were build from 1943 to 1949 and used until 1954. They had boilers for 10t/hr steam rate for 1500 ihp.
The exhaust turbine driving the fans had 200 horsepower, steam coming in at 1.6 atm and out at 1.02 atm.
The tenders had ten or twelve condensing units of 230 m² (2475 ft²) cooling surface each. The engines ran sucsessfully at 46 °C air temp. and gave ranges of
at least 1000 km (600 miles).
The much larger 4-8-4 SAR engines had boilers for around 3000 ihp.
The exhaust steam drove the draft turbine of 150 hp, then the fan turbine of 550 hp,
the back pressure being about the same as in a conventional loco (1 atm=15psi).
The five 7 ft- fans ran at 1000 rpm and blew 360 kg (800lb) per second , or
13,000,000 ft³ per hour of air through the fins to condense 24 tons/hr steam at
40 °C air temp.
One of the problems was the amount of lubricating oil in the feedwater. It was considered that 5 ppm or less was not dangerous. Sometimes peak amounts of
50 ppm were measured, although the 2-10-0s showed never more than 1-3 ppm.
After some experimenting it was found that the boilers tend to prime and the water mixed with the cylinder oil forming an emulsion. But the oil separators could only work if the oil was still a liquid and the steam a gas, the heavier oil drops were held back by a series of sheets, the steam streaming around them.
The solution was a "Vortex" water separator placed between the cylinders and
the oil separator. This let the oily water dripple to the ground, and the amount
of lub oil was never higher than 2-3 ppm thereafter.
Does anyone know how much oil a Stanley should have in the exhaust steam?
According to the figures I have, it should be 1000 ppm, that can`t be, can it ??
I know that the monotube boilers from the Doble-Henschel trucks and buses
were regularly sandblasted internally to clean them from carbon deposits.
Was or is this a usual practice in America ?
Martin</HTML>
Re: condensor/boiler area ratios
Posted by: George Nutz (IP Logged)
Date: September 13, 2002 06:24PM
<HTML>Martin,
So good to have you on the thread and great info on the German WW2 locomotives. Their condensing capabilities are impressive, wonder how they stack up to the performance of the Australian condensing Garrets? It appears that Abner recommended the sandblasting technique as the very end of the evaporation zone, with dry steam and tube temperatures climbing, would eventually clog up and overheat; the Achilles' hell of the high performance monotube. I understand some White monotube steam cars had the same problem eventually. Stanleys do not have the same problem tubewise but suffer greatly with crownsheet deposits. If a Stanley were to use one quart of oil per 500+ miles( full out about 320# steam/hr @ 35mph=about 5000# of steam developed per quart of oil) what would that be?? Very interesting that you have data on such stuff, does uni-essen place you in Germany? If so any information you could gather on the very large Sulzer monotube boilers would be of great interest, their control systems were an order of magnitude better than the Doble and were of a mechanical/proportional nature. I always wondered that after Warren Doble did all his work at Henschels and Borsig that the developement of the very large Sulzer monotubes took place.
Isn't this the greatest steam site!
Best, George Nutz</HTML>
So good to have you on the thread and great info on the German WW2 locomotives. Their condensing capabilities are impressive, wonder how they stack up to the performance of the Australian condensing Garrets? It appears that Abner recommended the sandblasting technique as the very end of the evaporation zone, with dry steam and tube temperatures climbing, would eventually clog up and overheat; the Achilles' hell of the high performance monotube. I understand some White monotube steam cars had the same problem eventually. Stanleys do not have the same problem tubewise but suffer greatly with crownsheet deposits. If a Stanley were to use one quart of oil per 500+ miles( full out about 320# steam/hr @ 35mph=about 5000# of steam developed per quart of oil) what would that be?? Very interesting that you have data on such stuff, does uni-essen place you in Germany? If so any information you could gather on the very large Sulzer monotube boilers would be of great interest, their control systems were an order of magnitude better than the Doble and were of a mechanical/proportional nature. I always wondered that after Warren Doble did all his work at Henschels and Borsig that the developement of the very large Sulzer monotubes took place.
Isn't this the greatest steam site!
Best, George Nutz</HTML>
Re: condensor/boiler area ratios
Posted by: tom ward (IP Logged)
Date: September 15, 2002 09:07AM
<HTML>Thax for the data on German locos. It appears that they used about 1.4 sq.ft. /lb./hr. It's the most direct reply I've had so far even though the diversions have been quite an education also. Never know where a basic question will take you. I've always claimed that genius is just the gift of asking better questions!</HTML>
Re: condensor/boiler area ratios
Posted by: Martin Werbeck (IP Logged)
Date: September 24, 2002 01:21PM
<HTML>Hello George !
According to your info`s, the Stanley needs about 1,7 lb oil per 5000 lb water,
or 340 lb oil per 1 mio. lb water, so the rate is 340 ppm.(parts per million).
Would be a bit difficult to get it down to 5 ppm ! The steam rate is quiet low at
3-5 lb/ft²hr, is this probably due to the burners low draft rather than the boiler design ?
Yes , you`re right, I´m studying here in Essen,Germany, and using the computer
of the Uni, because I can have a free search in the Internet. I have visited several
UK steam car tours and also the Great Dorset steam fair and had the great pleasure of riding in several steamers. I also had the great experience of driving
steam cars a few miles. It`s really amazing how fast a 30 HP Stanley can get away with a few turns of the engine !
When I look at your fantastic photo album, I am really astonished to see how many different models of steam cars still exist !
I have many of my informations from the British steam car club (Light Steam Power magazin) but also from the Uni`s library, especially from the VDI magazine (Verein Deutscher Ingenieure). I also found an interesting article and drawings about a Delling steam bus of 1929.
In the VDI is also the article about the Henschel-Doble engines from Richard Roosen. This has also been used in Walton`s Doble-book. Richard Roosen, who was the principal figure at Henschel`s for the condensing locos and the Doble- Henschel engines, also wrote most interesting book by himself.
In the VDI No 46 of 1933 is a full description and drawings of the Sulzer mototube boiler. It had a tube of 1.18" to 1.97" dia and 1300 metres (4300ft) total length, giving a heating surface of about 270 m²( 2900 ft²). Output was 7.5 t/hr at 100 atm
and up to 400 °C. The water was continually heated in the first 870 m (metres) of
tube of convective surface, then flashed to steam in the next 180 m , at this point additional water was injected, then the steam was heated to its final temperature in the last 220 m. About 72 m² of the surface was in the combustion chamber.
At 7,28 t/hr total steam rate, 1,97 t/hr were injected by the normaliser to get 370 °C end temperature.
The velocity in the tube varied from 1,5 to 3 m/sec in the economizer to 18 to 25 m/sec in the superheater. A drop of water would therefore need 8 2/3 min to go through the tube, but as much as 17 min when working at half output. Any increase in feedwater amount/pressure would go through the tube in less than 1 second, (because water is not compressible) up to the point were it would flash into steam in the hot tube. The additional steam would than need only 37 seconds to the end of the tube at full load.
Realising that the different time lags were the main reason for temperature fluctuation, the Sulzer engineers designed a quiet complicated hydraulic (oil) system connected to a number of thermostats.
A "Differentialregler" controlled the amount of feedwater in reaction to the time lag of the increase in tempetature at different points in the tube, and also in a direct way from the final temperature. The pressure in the hydraulic system was therefore influenced both by the superheat temperature and indirect by the velocity at which the water/steam went through the tube and caused a different fast increase/decrease in temperature at different points of the tube.
Being designed for stationary use it showed little difference in superheat when the pressure was dropped from about 90 to 60 atm in 3 minutes. But this would probably be to sluggish for automobile use with much more rapid load changes.
I find it next to impossible to decribe the whole system whithout using diagrams and drawings, but I hope I could help you a little bit !
With best regards, Martin</HTML>
According to your info`s, the Stanley needs about 1,7 lb oil per 5000 lb water,
or 340 lb oil per 1 mio. lb water, so the rate is 340 ppm.(parts per million).
Would be a bit difficult to get it down to 5 ppm ! The steam rate is quiet low at
3-5 lb/ft²hr, is this probably due to the burners low draft rather than the boiler design ?
Yes , you`re right, I´m studying here in Essen,Germany, and using the computer
of the Uni, because I can have a free search in the Internet. I have visited several
UK steam car tours and also the Great Dorset steam fair and had the great pleasure of riding in several steamers. I also had the great experience of driving
steam cars a few miles. It`s really amazing how fast a 30 HP Stanley can get away with a few turns of the engine !
When I look at your fantastic photo album, I am really astonished to see how many different models of steam cars still exist !
I have many of my informations from the British steam car club (Light Steam Power magazin) but also from the Uni`s library, especially from the VDI magazine (Verein Deutscher Ingenieure). I also found an interesting article and drawings about a Delling steam bus of 1929.
In the VDI is also the article about the Henschel-Doble engines from Richard Roosen. This has also been used in Walton`s Doble-book. Richard Roosen, who was the principal figure at Henschel`s for the condensing locos and the Doble- Henschel engines, also wrote most interesting book by himself.
In the VDI No 46 of 1933 is a full description and drawings of the Sulzer mototube boiler. It had a tube of 1.18" to 1.97" dia and 1300 metres (4300ft) total length, giving a heating surface of about 270 m²( 2900 ft²). Output was 7.5 t/hr at 100 atm
and up to 400 °C. The water was continually heated in the first 870 m (metres) of
tube of convective surface, then flashed to steam in the next 180 m , at this point additional water was injected, then the steam was heated to its final temperature in the last 220 m. About 72 m² of the surface was in the combustion chamber.
At 7,28 t/hr total steam rate, 1,97 t/hr were injected by the normaliser to get 370 °C end temperature.
The velocity in the tube varied from 1,5 to 3 m/sec in the economizer to 18 to 25 m/sec in the superheater. A drop of water would therefore need 8 2/3 min to go through the tube, but as much as 17 min when working at half output. Any increase in feedwater amount/pressure would go through the tube in less than 1 second, (because water is not compressible) up to the point were it would flash into steam in the hot tube. The additional steam would than need only 37 seconds to the end of the tube at full load.
Realising that the different time lags were the main reason for temperature fluctuation, the Sulzer engineers designed a quiet complicated hydraulic (oil) system connected to a number of thermostats.
A "Differentialregler" controlled the amount of feedwater in reaction to the time lag of the increase in tempetature at different points in the tube, and also in a direct way from the final temperature. The pressure in the hydraulic system was therefore influenced both by the superheat temperature and indirect by the velocity at which the water/steam went through the tube and caused a different fast increase/decrease in temperature at different points of the tube.
Being designed for stationary use it showed little difference in superheat when the pressure was dropped from about 90 to 60 atm in 3 minutes. But this would probably be to sluggish for automobile use with much more rapid load changes.
I find it next to impossible to decribe the whole system whithout using diagrams and drawings, but I hope I could help you a little bit !
With best regards, Martin</HTML>
Re: condensor/boiler area ratios
Posted by: Peter Brow (IP Logged)
Date: October 08, 2002 07:02AM
<HTML>What kind of internal/external square footage is obtained per square foot of typical modern radiator core, assuming 1" core thickness? Would increasing the internal square footage improve condensing more than increasing external square footage by the same amount?
I ask because I noticed that some of the radiators for which better condensing was claimed, like the Doble-Detroit honeycomb radiator and the Saab V-front minitubular radiator, seem to have larger internal surface area per unit of external surface area. This _seems_ to suggest that there may be an advantage to increasing internal (steamed) surface area.
But maybe current production radiators have achieved the same (or better) heat-exchange improvement through increased fin density?
Peter</HTML>
I ask because I noticed that some of the radiators for which better condensing was claimed, like the Doble-Detroit honeycomb radiator and the Saab V-front minitubular radiator, seem to have larger internal surface area per unit of external surface area. This _seems_ to suggest that there may be an advantage to increasing internal (steamed) surface area.
But maybe current production radiators have achieved the same (or better) heat-exchange improvement through increased fin density?
Peter</HTML>
Re: condensor/boiler area ratios
Posted by: Peter Heid (IP Logged)
Date: October 09, 2002 12:49PM
<HTML>Peter,
I believe the more steam side surface area the better as long as you can get air to flow through it. To condense larger amounts of steam requires a larger steam side surface area or a more efficient condenser. Assuming no changes can be made on the steam side, to increase the capacity of a condensor of a specific size, the air side must become more efficient. More surface area in the form of fins can help but to a limited extent because they tend to block the air flow as fin density increases. The Vee front radiator is said to pass 4 times the air as a conventional radiator. At speed, the flat radiator has a great deal of air that bounces off and passes around it instead of passing through even with a fan on the back side. If you measured the pressure in the Vee of the Vee front radiator, you would notice a drop in pressure as you aproach the apex to a point where debris and bugs fall down the Vee instead of being forced theough. The air passage through the fins in the Vee is slowed because of the greater surface area of the Vee configuration and fin spacing is not used to control the speed of the air flow, to as great an extent as is done in a conventional radiator. Of course if the air or steam flows through too fast the opportunity to exchange heat will be lost.
Peter Heid</HTML>
I believe the more steam side surface area the better as long as you can get air to flow through it. To condense larger amounts of steam requires a larger steam side surface area or a more efficient condenser. Assuming no changes can be made on the steam side, to increase the capacity of a condensor of a specific size, the air side must become more efficient. More surface area in the form of fins can help but to a limited extent because they tend to block the air flow as fin density increases. The Vee front radiator is said to pass 4 times the air as a conventional radiator. At speed, the flat radiator has a great deal of air that bounces off and passes around it instead of passing through even with a fan on the back side. If you measured the pressure in the Vee of the Vee front radiator, you would notice a drop in pressure as you aproach the apex to a point where debris and bugs fall down the Vee instead of being forced theough. The air passage through the fins in the Vee is slowed because of the greater surface area of the Vee configuration and fin spacing is not used to control the speed of the air flow, to as great an extent as is done in a conventional radiator. Of course if the air or steam flows through too fast the opportunity to exchange heat will be lost.
Peter Heid</HTML>
Re: condensor/boiler area ratios
Posted by: Peter Brow (IP Logged)
Date: October 24, 2002 05:36AM
<HTML>Hi Peter,
Thanks; did some thinkwork & sketching on the Saab vee-front radiator tonight, and noticed that the heat-exchange surface area per unit of radiator volume is less, due to the open spaces in the vees. Also, the Saab drawings don't show any fins, just tubes. However, it would be easy to design this thing for much better airflow than a conventional radiator, practically zero flowpath restriction if desired (though there will be turbulence as air passes around the tubes). IE, the area of airflow opening between tubes can be made the same as the frontal area of the radiator, or more if desired. The principle is similar to the corrugated surfaces of low-restriction air, fuel, & oil filters in gas cars.
Also, I think that another part of the heat-exchange improvement in the vee radiator comes from the fact that every square inch of the tubes receives cool ambient air rather than air heated by passing over hot surfaces in front of it. This means a higher temperature differential and therefore higher heat transfer per unit area, relative to a conventional tube/fin radiator. In fact, every square inch of the heat-exchange tubing has the maximum possible temperature differential. This occurred to me when considering Ron's note (elsewhere on this forum) that making a conventional radiator core thicker gives diminishing returns due to the increasingly hotter air flowing over any extra heat-exchange surface area added to the back of the radiator. The wizards of Trollhattan eliminated this problem in their steam "bil".
For those not familiar with the Saab "vee-front" radiator, this is not a conventional radiator with a vee shape, like in 1915-1918 Stanleys, but rather a flat-fronted radiator core with two rows of small widely-spaced vertical header tubes, one front (inlet) and one rear (outlet). The two rows are staggered/offset relative to each other, and each front header tube has two angled columns of horizontal cooling tubes running back to the two closest rear header tubes, in "vee" shapes. A horizontal cross-section view (looking down) shows a zig-zag of horizontal tubes, with front & rear vertical tubes as the "points" of each vee-shaped zig (or zag).
Hope that description is not too convoluted to visualize. If anyone is interested yet still perplexed, Maimonides here will mail you a copy of the drawings. I plan to scan this thing and get it on one of my steam webpages in the near future. It is an intriguing design lifted from the Philips Stirling engine coolers. I'm not sure if Ford used this type of radiator in their experimental circa 1975 Stirling-engined Grand Torino, but I doubt it. That thing was a pooch. Real "disproof of concept" stuff. Most efficient small-engine type in the world mated to an early-'70s Ford slush box, whew. The design team must have used crayons.
Peter B.</HTML>
Thanks; did some thinkwork & sketching on the Saab vee-front radiator tonight, and noticed that the heat-exchange surface area per unit of radiator volume is less, due to the open spaces in the vees. Also, the Saab drawings don't show any fins, just tubes. However, it would be easy to design this thing for much better airflow than a conventional radiator, practically zero flowpath restriction if desired (though there will be turbulence as air passes around the tubes). IE, the area of airflow opening between tubes can be made the same as the frontal area of the radiator, or more if desired. The principle is similar to the corrugated surfaces of low-restriction air, fuel, & oil filters in gas cars.
Also, I think that another part of the heat-exchange improvement in the vee radiator comes from the fact that every square inch of the tubes receives cool ambient air rather than air heated by passing over hot surfaces in front of it. This means a higher temperature differential and therefore higher heat transfer per unit area, relative to a conventional tube/fin radiator. In fact, every square inch of the heat-exchange tubing has the maximum possible temperature differential. This occurred to me when considering Ron's note (elsewhere on this forum) that making a conventional radiator core thicker gives diminishing returns due to the increasingly hotter air flowing over any extra heat-exchange surface area added to the back of the radiator. The wizards of Trollhattan eliminated this problem in their steam "bil".
For those not familiar with the Saab "vee-front" radiator, this is not a conventional radiator with a vee shape, like in 1915-1918 Stanleys, but rather a flat-fronted radiator core with two rows of small widely-spaced vertical header tubes, one front (inlet) and one rear (outlet). The two rows are staggered/offset relative to each other, and each front header tube has two angled columns of horizontal cooling tubes running back to the two closest rear header tubes, in "vee" shapes. A horizontal cross-section view (looking down) shows a zig-zag of horizontal tubes, with front & rear vertical tubes as the "points" of each vee-shaped zig (or zag).
Hope that description is not too convoluted to visualize. If anyone is interested yet still perplexed, Maimonides here will mail you a copy of the drawings. I plan to scan this thing and get it on one of my steam webpages in the near future. It is an intriguing design lifted from the Philips Stirling engine coolers. I'm not sure if Ford used this type of radiator in their experimental circa 1975 Stirling-engined Grand Torino, but I doubt it. That thing was a pooch. Real "disproof of concept" stuff. Most efficient small-engine type in the world mated to an early-'70s Ford slush box, whew. The design team must have used crayons.
Peter B.</HTML>
Re: condensor/boiler area ratios
Posted by: george sosin (IP Logged)
Date: March 30, 2004 03:21PM
<HTML>
Hello,
I am a teacher of English in a French Institute of Technology dealing with Heating and Air Conditioning.
Could you please give me some information about web sites where I could possibly find more information about the above subjects as well as British companies dealing with heating installations.
Thank you in advance</HTML>
Hello,
I am a teacher of English in a French Institute of Technology dealing with Heating and Air Conditioning.
Could you please give me some information about web sites where I could possibly find more information about the above subjects as well as British companies dealing with heating installations.
Thank you in advance</HTML>
Re: condensor design
Posted by: Ichsan (IP Logged)
Date: August 24, 2004 08:30AM
<HTML>Dear there,
My name is Ichsan, and I've college at Pasundan University. I'm interest about re-condensor design that you've made. I hope you can send to my e mail about :
- How to re-design a condensor for cooling system ?
- How to build it ?
- Can you send me the schematic of the system ?
Thank's a lot
Ichsan</HTML>
My name is Ichsan, and I've college at Pasundan University. I'm interest about re-condensor design that you've made. I hope you can send to my e mail about :
- How to re-design a condensor for cooling system ?
- How to build it ?
- Can you send me the schematic of the system ?
Thank's a lot
Ichsan</HTML>
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