Williams ws. Rankin
Posted by Howard Langdon
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Re: Williams ws. Rankin July 22, 2006 08:18PM |
Registered: 19 years ago Posts: 318 |
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Re: Williams ws. Rankin July 23, 2006 10:51AM |
Registered: 19 years ago Posts: 1,441 |
Hi Howard,
OK, are we talking about the original Series D, E and F Doble temperature control systems? Two pumps-four pumps with a normalizer, and the two speed steam pump water feed and normalizer in the Series F cars.
If you are asking about the triple tube and the quad tube thermostats, they were abject failures. This was the period when Abner and Warren were desparately looking for a control system that would work perfectly with steam generators that were being forced to the limit, around 28 lbs/hr/sq/ft. The notebooks and the letters between the brothers are simply full of dozens of schemes, none of which worked at all, or as well, as the single tube quartz thermostat.
Having cold feedwater in one tube, saturated steam in another, and dry steam in another and superheated in the last one, guanteed that the thermostat assembly bent and warped like a spring. Not to forget the awful number of connections that had to be made in the steam generator into and out of these thermostats.
Hot on one side and cold on the other, with the tubes welded to the quartz rod tube was not the brightest of Abner's ideas. Eventually this broke the quartz rods putting the thermostat out of action.
In the engineering notebooks, and the Sentinel and Price notes, he designed seven of them before giving up. Note that in the last Doble system, the McCulloch car, Abner went back to the single tube thermostat.
A good Doble quartz rod thermostat can easily hold temperature to about 2°F. BUT; getting one to do that means months of most careful adjustment, lot of test driving, and knowing whether to make corrections on the control box main temperature control, or with the compensator.
Also, adjusting the normalizer oriface to sqirt in just the right amount of water. A very precise adjustment.
It takes a most delicate adjustment between the three before the car really settles down and runs as it should.
Also, we have found some cars with very short thermostats, and not the longer version that they first put in the Dobles. Previous modifications by owners who really didn't understand what was needed.
These make VERY little movement as compared to the longer ones and are not as sensitive. I much prefer to put in long quartz rod assemblies, because they quickly give more motion with temperature change and are easier to set precisely.
There is one other problem with making Doble style quartz rod thermostats. The fit between the rods and the thin tube they ride in, has to be a rather loose fit. The tubes sometimes warp or bend over time, because some people don't put in the pin on the end that supports the whole assembly; but just leave it hanging there. When the rods crack, or fret on the ends, then the whole thing goes out of whack and will not control correctly. As many as possible clear across the superheater coil and supported at the end.
Once done correctly, a Doble quartz rod thermostat will work just fine for many years and will be very stable and quick acting.
Quite honestly, I like to use dual thermocouples, one on the outlet of the superheater and one just before the normalizer inlet. Simple purchased little control modules and they run the big power relay and blower motor and the pump solenoid valves. Can be set very precisely and they work just fine.
However, when one is restoring a Doble, only the quartz rod thermostats are used, because I don't like modifying antique steamers. New, yes, old, no.
JC
Edited 2 time(s). Last edit at 07/24/2006 10:13AM by James D. Crank.
OK, are we talking about the original Series D, E and F Doble temperature control systems? Two pumps-four pumps with a normalizer, and the two speed steam pump water feed and normalizer in the Series F cars.
If you are asking about the triple tube and the quad tube thermostats, they were abject failures. This was the period when Abner and Warren were desparately looking for a control system that would work perfectly with steam generators that were being forced to the limit, around 28 lbs/hr/sq/ft. The notebooks and the letters between the brothers are simply full of dozens of schemes, none of which worked at all, or as well, as the single tube quartz thermostat.
Having cold feedwater in one tube, saturated steam in another, and dry steam in another and superheated in the last one, guanteed that the thermostat assembly bent and warped like a spring. Not to forget the awful number of connections that had to be made in the steam generator into and out of these thermostats.
Hot on one side and cold on the other, with the tubes welded to the quartz rod tube was not the brightest of Abner's ideas. Eventually this broke the quartz rods putting the thermostat out of action.
In the engineering notebooks, and the Sentinel and Price notes, he designed seven of them before giving up. Note that in the last Doble system, the McCulloch car, Abner went back to the single tube thermostat.
A good Doble quartz rod thermostat can easily hold temperature to about 2°F. BUT; getting one to do that means months of most careful adjustment, lot of test driving, and knowing whether to make corrections on the control box main temperature control, or with the compensator.
Also, adjusting the normalizer oriface to sqirt in just the right amount of water. A very precise adjustment.
It takes a most delicate adjustment between the three before the car really settles down and runs as it should.
Also, we have found some cars with very short thermostats, and not the longer version that they first put in the Dobles. Previous modifications by owners who really didn't understand what was needed.
These make VERY little movement as compared to the longer ones and are not as sensitive. I much prefer to put in long quartz rod assemblies, because they quickly give more motion with temperature change and are easier to set precisely.
There is one other problem with making Doble style quartz rod thermostats. The fit between the rods and the thin tube they ride in, has to be a rather loose fit. The tubes sometimes warp or bend over time, because some people don't put in the pin on the end that supports the whole assembly; but just leave it hanging there. When the rods crack, or fret on the ends, then the whole thing goes out of whack and will not control correctly. As many as possible clear across the superheater coil and supported at the end.
Once done correctly, a Doble quartz rod thermostat will work just fine for many years and will be very stable and quick acting.
Quite honestly, I like to use dual thermocouples, one on the outlet of the superheater and one just before the normalizer inlet. Simple purchased little control modules and they run the big power relay and blower motor and the pump solenoid valves. Can be set very precisely and they work just fine.
However, when one is restoring a Doble, only the quartz rod thermostats are used, because I don't like modifying antique steamers. New, yes, old, no.
JC
Edited 2 time(s). Last edit at 07/24/2006 10:13AM by James D. Crank.
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Re: Williams ws. Rankin July 24, 2006 07:54PM |
Registered: 19 years ago Posts: 318 |
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Re: Williams ws. Rankin July 24, 2006 11:19PM |
Registered: 19 years ago Posts: 1,441 |
Howard,
Yes, when set up right, the two temperature contacts just right, second set of pumps ON and the normalizer coming into action when they should, and the normalizer oriface correct, you are hard pressed to see the outlet temperature gauge move at all. E-14 moves about one half the needle width under the hardest driving.
This takes a lot of time driving and adjusting the contacts to get a Doble to do this; but the good ones do it all day.
Pressure is by a separate contact in the control box that is operated by a spring loaded diaphram. Same as the Stanley; but with a contact in place of their fuel valve.
JC
Yes, when set up right, the two temperature contacts just right, second set of pumps ON and the normalizer coming into action when they should, and the normalizer oriface correct, you are hard pressed to see the outlet temperature gauge move at all. E-14 moves about one half the needle width under the hardest driving.
This takes a lot of time driving and adjusting the contacts to get a Doble to do this; but the good ones do it all day.
Pressure is by a separate contact in the control box that is operated by a spring loaded diaphram. Same as the Stanley; but with a contact in place of their fuel valve.
JC
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Re: Williams ws. Rankin July 25, 2006 07:58PM |
Registered: 19 years ago Posts: 318 |
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Re: Williams ws. Rankin September 17, 2006 07:24PM |
Registered: 19 years ago Posts: 318 |
HI Andy.derr boilers had two rose of tubes under the super heeter.so vary had to have alert tubing to get not much super heat I don’t thing it will ovary much over 650 f.which is to loge for a stanly.any marten engine vat wont run 1000 f+ have batter stay onpaper don’t worry about the oil.dzine it so the oil wont get to hot like the Williams did. At 1000(PSI) 1000(F) at vees steam conditions you have about 500(Btu) per pound to do work. You need hi super heat
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Re: Williams ws. Rankin September 18, 2006 04:35PM |
Registered: 19 years ago Posts: 1,038 |
Hi Howard.
Analysis of lower pressure 500 PSIA engines with little compression and long cutoff to higher pressure 1000 PSIA with short cutoff and compression shows supper heat produces less efficiency increases in the high pressure short cutoff engines.
I did two cycles for 500 PSIA inlet expanding to 300 PSIA and exhausting to 15 PSIA.
Going from 600F inlet temperature to 1000F inlet pressure produces an efficiency increase 127% 65.4 and 65.7 percent cutoff
Increassing expansion to 18 PSIA 10 and 12 % cutoff the efficiency increase, 600F to 1000F inlet, drops to 118%
Two cases: 1000 PSIA inlet engine expanding to 18 PSIA.
With compression to 50 shows an even less gain, 600F to 1000F inlet, of 116%
With full expansion the gain drops, 600F to 1000F inlet, to 112%
As you go to higher expansions with compression eliminating clearance losses the gain of super heating decreases. It takes a lot more super heat to effect efficiency. Once exhaust is dry steam the gain / degree super heat drops.
It looks like the more efficient the engine the less gain there is from super heating. As long as you have enough to have close to dry steam exhaust.
Breaking the cycle into parts I see larger super heating efficiency gains comming from the non-expansive admittance process. The expansion part has much smaller relatice gaines.
A 66% cutoff engine should show a much greater gane in efficiency with higher super heat then a 5% cutoff engine.
I think Jerry posted something about over super heating a while back.
Reading my on post. It wasn't real clear.
Also was using same end expansion pressure not expansion ratio.
As it turns out using a fixed end of expansion instead of ratio make a lot of differance. Expansion to a fixed pressure is what gave super heat the effect I described. As you increase super heat it takes less volmetric expansion to effect the same pressure drop. It is the lower expansion ratio that is canceling efficiency gain of expansion.
Andy
Edited 1 time(s). Last edit at 09/19/2006 12:22PM by Andy.
Analysis of lower pressure 500 PSIA engines with little compression and long cutoff to higher pressure 1000 PSIA with short cutoff and compression shows supper heat produces less efficiency increases in the high pressure short cutoff engines.
I did two cycles for 500 PSIA inlet expanding to 300 PSIA and exhausting to 15 PSIA.
Going from 600F inlet temperature to 1000F inlet pressure produces an efficiency increase 127% 65.4 and 65.7 percent cutoff
Increassing expansion to 18 PSIA 10 and 12 % cutoff the efficiency increase, 600F to 1000F inlet, drops to 118%
Two cases: 1000 PSIA inlet engine expanding to 18 PSIA.
With compression to 50 shows an even less gain, 600F to 1000F inlet, of 116%
With full expansion the gain drops, 600F to 1000F inlet, to 112%
As you go to higher expansions with compression eliminating clearance losses the gain of super heating decreases. It takes a lot more super heat to effect efficiency. Once exhaust is dry steam the gain / degree super heat drops.
It looks like the more efficient the engine the less gain there is from super heating. As long as you have enough to have close to dry steam exhaust.
Breaking the cycle into parts I see larger super heating efficiency gains comming from the non-expansive admittance process. The expansion part has much smaller relatice gaines.
A 66% cutoff engine should show a much greater gane in efficiency with higher super heat then a 5% cutoff engine.
I think Jerry posted something about over super heating a while back.
Reading my on post. It wasn't real clear.
Also was using same end expansion pressure not expansion ratio.
As it turns out using a fixed end of expansion instead of ratio make a lot of differance. Expansion to a fixed pressure is what gave super heat the effect I described. As you increase super heat it takes less volmetric expansion to effect the same pressure drop. It is the lower expansion ratio that is canceling efficiency gain of expansion.
Andy
Edited 1 time(s). Last edit at 09/19/2006 12:22PM by Andy.
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Re: Williams ws. Rankin September 19, 2006 08:38PM |
Registered: 19 years ago Posts: 318 |
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Re: Williams ws. Rankin September 20, 2006 11:50AM |
Registered: 19 years ago Posts: 1,038 |
Howard.
Howard you are partly correct. It is a heat engine and the more heat you put in the more work you can extract. So you are absoultly right. The key there, is extraction. Puting more heat in and not geting more heat out is a waste of that heat. You can't just look at one paremeter, like super heat, and make a blanket statement. There are a lot of inter-related process going on in a steam engine. The exhaust pressure puts a limit on the expansion ratio. Expanding below exhaust pressure puts efficiency into the shitter.
Lets look at an expansion example. Say you havve 1000 PSIA inlet pressure and 15 PSIA exhaust back pressure. Basicly atmospheric exhaust pressure.
Lets set the expansion ratio at 33.33:1 and see how inlet teperature effects the end of expansion pressure.
Temp pressure
600F 18.5 PSIA
700F 16.4 PSIA
800F 15.0 PSIA
900F 13.9 PSIA
1000F 13.0 PSIA
1100F 12.3 PSIA
1200F 11.7 PSIA
So as we increase the steam temperature the end of expansion pressure decreases. And we are looking at aproxmatly an atmospheric exhaust pressure of 15 PSIA. At 900F we are into over expansion. How do you think the engine will run in that state.
To avoid over expansion. What I did was to use a set end of expansion pressure just slightly above exhaust pressure. So basicly the max expansion between max inlet pressure and exhaust pressure. More expansion would put the end of expansion below exhaust pressure.
I looked at the admittance(constant pressure process) and expansion(isentropic processes seperatly). I ploted the work produced by each process and the total work divided by the heat in. Basicly efficiency of each process and engine.
What it shows is that the constant pressure process made great gains from supper heat. You can increase the efficiency 150%. Wow!! 150% doesn't that sound great. But going from 8% efficiency to 12% efficiency is really nothing to brag about.
Looking at the expansion process as another story. With a fixed expansion ratio. It also gains efficiency. But the rub comes in when you are allready expanding close to exhaust pressure. Increassing heat content increases the pressure and temperature drop. So you would expect an efficiecy gain. But over expanding below exhaust pressure means exhaust flows into the cylander. Efficincy goes into the shiter. Fixing the exhaust pressure, by varing the expansion ratio, means that as you super heat you also reduce the expansion ratio to keep from over expanding. Reducing expansion reduces the temperature and pressure drop and less work is extracted by the expansion process. In fact the plot shows a initially reduction in efficiency of the expansion process. The net result is that at close to full expansion you get less efficiency increase per BTU of super heat.
Going from 600F to 1200F steam temp: A non-expanding engine can get 150% efficiency gain. 8% to 12%, 4 percentage points.
A high expansion engine 1000 PSIA expanding to 15 PSIA gets around 2.4 percentage points gain. 32% to 34.4% efficiency gain isn't much for the high cost of higher temperature.
To be fair. This probably doesn't apply to anything your doing. My engine design is expanding close to exhaust pressure and analysys showed this problem.
Edited 2 time(s). Last edit at 09/20/2006 12:07PM by Andy.
Howard you are partly correct. It is a heat engine and the more heat you put in the more work you can extract. So you are absoultly right. The key there, is extraction. Puting more heat in and not geting more heat out is a waste of that heat. You can't just look at one paremeter, like super heat, and make a blanket statement. There are a lot of inter-related process going on in a steam engine. The exhaust pressure puts a limit on the expansion ratio. Expanding below exhaust pressure puts efficiency into the shitter.
Lets look at an expansion example. Say you havve 1000 PSIA inlet pressure and 15 PSIA exhaust back pressure. Basicly atmospheric exhaust pressure.
Lets set the expansion ratio at 33.33:1 and see how inlet teperature effects the end of expansion pressure.
Temp pressure
600F 18.5 PSIA
700F 16.4 PSIA
800F 15.0 PSIA
900F 13.9 PSIA
1000F 13.0 PSIA
1100F 12.3 PSIA
1200F 11.7 PSIA
So as we increase the steam temperature the end of expansion pressure decreases. And we are looking at aproxmatly an atmospheric exhaust pressure of 15 PSIA. At 900F we are into over expansion. How do you think the engine will run in that state.
To avoid over expansion. What I did was to use a set end of expansion pressure just slightly above exhaust pressure. So basicly the max expansion between max inlet pressure and exhaust pressure. More expansion would put the end of expansion below exhaust pressure.
I looked at the admittance(constant pressure process) and expansion(isentropic processes seperatly). I ploted the work produced by each process and the total work divided by the heat in. Basicly efficiency of each process and engine.
What it shows is that the constant pressure process made great gains from supper heat. You can increase the efficiency 150%. Wow!! 150% doesn't that sound great. But going from 8% efficiency to 12% efficiency is really nothing to brag about.
Looking at the expansion process as another story. With a fixed expansion ratio. It also gains efficiency. But the rub comes in when you are allready expanding close to exhaust pressure. Increassing heat content increases the pressure and temperature drop. So you would expect an efficiecy gain. But over expanding below exhaust pressure means exhaust flows into the cylander. Efficincy goes into the shiter. Fixing the exhaust pressure, by varing the expansion ratio, means that as you super heat you also reduce the expansion ratio to keep from over expanding. Reducing expansion reduces the temperature and pressure drop and less work is extracted by the expansion process. In fact the plot shows a initially reduction in efficiency of the expansion process. The net result is that at close to full expansion you get less efficiency increase per BTU of super heat.
Going from 600F to 1200F steam temp: A non-expanding engine can get 150% efficiency gain. 8% to 12%, 4 percentage points.
A high expansion engine 1000 PSIA expanding to 15 PSIA gets around 2.4 percentage points gain. 32% to 34.4% efficiency gain isn't much for the high cost of higher temperature.
To be fair. This probably doesn't apply to anything your doing. My engine design is expanding close to exhaust pressure and analysys showed this problem.
Edited 2 time(s). Last edit at 09/20/2006 12:07PM by Andy.
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Re: Williams ws. Rankin September 23, 2006 09:16PM |
Registered: 19 years ago Posts: 318 |
HI Andy now we get to sum thing I have bin going to post. To many expanses is worst van not enough. As you say.bot are you zinging twice as many as work in the reel willed. The f doble 3 by 5.2 by 5 short cuts off was 40% in cluing clearance. It has 5 extspansuns.exzort tenp is 250 f the in let temp is 850 f ps I have bin off line for 3 days
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Re: Williams ws. Rankin September 28, 2006 03:03PM |
Registered: 19 years ago Posts: 1,038 |
Hi Howard.
Not quite sure what you’re trying to say about the Doble there. What’s the point? The ending temp or what.
I have said many times over, expansion is bad on efficiency. Jerry has also said the same thing. Nothing new.
For an automobile, power (or torque) range is an important design issue. How you control power is important. Realizing that over expansion drastically hurts efficiency and the bottom line cost of operation. It’s important to consider avoiding over expansion or at least limiting it.
What I have been doing is not a simple cycle analysis. A steam engine doesn't operate on just one cycle. It operates over a power range of many different cycles having different throttling, cutoffs, and etc.
We can set some design goals for speed, acceleration and so forth. Speed is of primary importance. A design goal of a top speed of say 100 MPH sets a speed range requirement. Taking clues from current cars I set the lowest speed to be 5 MPH. That is what several cars I have checked, having automatic transmissions, do at an idle. Below that breaking is required. 5 MPH to 100 MPH is a 20:1 speed range. Requiring a 400:1 torque range. A 8000:1 power range.
Getting good efficiency over such a huge power range is complicated. I think it can be done and would have a completive MPG rating. But there are some trade offs in the design that need investigating. An example would be number of cylinders. More cylinders would allow shorter cutoff at low speed and starting. Working with a specific displacement requirement for the speed and acceleration performance goals. More cylinders equals smaller displacements per cylinder and a higher surface area to volume ratio. We get low speed efficiency gains from shorter cutoff and potentially higher heat loss from more surface area. We can counter heat loss with surface coatings to some extent at a cost$. There is no quantitative data available that can answer how many cylinders is best.
In an automobile we are interested in MPG. Fuel use is proportional to power/efficiency. Power being proportional to the speed cubed makes MPG proportional to efficiency/speed^3. Efficiency at normal cursing speed is more important then at minim speed. It’s even a bit more complicated because the cubed rule applies to turbulent flow wind resistance. At low speeds laminar flow would apply and be a square rule. The speed at which turbulent flow applies or laminar flow applies is very dependent on body design.
Super-heat also effects power range when throttling. Increased super-heat generally means more pressure drop during expansion for a given expansion. Your efficient range of throttling is from max inlet pressure to the point where over expansion occurs. It is the end of expansion pressure and exhaust pressure that sets the efficient throttling range. Dropping the end of expansion pressure decreases the throttling range or goes into over expansion at the low end.
At any rate I think you get what I was saying about super heat. It’s one parameter of many that must all fit together. You can have to much or to little of any one thing. I think you need to design the engine for the vehicle and its intended use.
Andy
Not quite sure what you’re trying to say about the Doble there. What’s the point? The ending temp or what.
I have said many times over, expansion is bad on efficiency. Jerry has also said the same thing. Nothing new.
For an automobile, power (or torque) range is an important design issue. How you control power is important. Realizing that over expansion drastically hurts efficiency and the bottom line cost of operation. It’s important to consider avoiding over expansion or at least limiting it.
What I have been doing is not a simple cycle analysis. A steam engine doesn't operate on just one cycle. It operates over a power range of many different cycles having different throttling, cutoffs, and etc.
We can set some design goals for speed, acceleration and so forth. Speed is of primary importance. A design goal of a top speed of say 100 MPH sets a speed range requirement. Taking clues from current cars I set the lowest speed to be 5 MPH. That is what several cars I have checked, having automatic transmissions, do at an idle. Below that breaking is required. 5 MPH to 100 MPH is a 20:1 speed range. Requiring a 400:1 torque range. A 8000:1 power range.
Getting good efficiency over such a huge power range is complicated. I think it can be done and would have a completive MPG rating. But there are some trade offs in the design that need investigating. An example would be number of cylinders. More cylinders would allow shorter cutoff at low speed and starting. Working with a specific displacement requirement for the speed and acceleration performance goals. More cylinders equals smaller displacements per cylinder and a higher surface area to volume ratio. We get low speed efficiency gains from shorter cutoff and potentially higher heat loss from more surface area. We can counter heat loss with surface coatings to some extent at a cost$. There is no quantitative data available that can answer how many cylinders is best.
In an automobile we are interested in MPG. Fuel use is proportional to power/efficiency. Power being proportional to the speed cubed makes MPG proportional to efficiency/speed^3. Efficiency at normal cursing speed is more important then at minim speed. It’s even a bit more complicated because the cubed rule applies to turbulent flow wind resistance. At low speeds laminar flow would apply and be a square rule. The speed at which turbulent flow applies or laminar flow applies is very dependent on body design.
Super-heat also effects power range when throttling. Increased super-heat generally means more pressure drop during expansion for a given expansion. Your efficient range of throttling is from max inlet pressure to the point where over expansion occurs. It is the end of expansion pressure and exhaust pressure that sets the efficient throttling range. Dropping the end of expansion pressure decreases the throttling range or goes into over expansion at the low end.
At any rate I think you get what I was saying about super heat. It’s one parameter of many that must all fit together. You can have to much or to little of any one thing. I think you need to design the engine for the vehicle and its intended use.
Andy
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Re: Williams ws. Rankin September 28, 2006 03:44PM |
Registered: 19 years ago Posts: 835 |
H Andy
I can't agree with you more. This is what I was talking about a power band. A long stroke slow speed engine has none. We are throttling with cutoff. At low speed there is a small engine, at high it is large through RPM. Six cyl was the choice to minimize cutoff. Vary clearance with RPM, constant temp and preasure.
Supercritical allows for a small valve because of short cutoff. Every thing gets smaller.
Harry
I can't agree with you more. This is what I was talking about a power band. A long stroke slow speed engine has none. We are throttling with cutoff. At low speed there is a small engine, at high it is large through RPM. Six cyl was the choice to minimize cutoff. Vary clearance with RPM, constant temp and preasure.
Supercritical allows for a small valve because of short cutoff. Every thing gets smaller.
Harry
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Re: Williams ws. Rankin September 28, 2006 04:38PM |
Registered: 19 years ago Posts: 1,038 |
Hi Harry
That's not exactly what I had in mind. But it works. Maybe? But I still doubt high RPM to work all that well. Havn't seen anything that proves the torque dropoff above 1000 RPM, seen in every piston steam engines dyno run published, to be any different in reciently developed high RPM engines. Jay Carter's engine or yours. I hope you have solved that problem. How about a dyno run?
Andy
That's not exactly what I had in mind. But it works. Maybe? But I still doubt high RPM to work all that well. Havn't seen anything that proves the torque dropoff above 1000 RPM, seen in every piston steam engines dyno run published, to be any different in reciently developed high RPM engines. Jay Carter's engine or yours. I hope you have solved that problem. How about a dyno run?
Andy
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Re: Williams ws. Rankin September 28, 2006 05:10PM |
Registered: 19 years ago Posts: 835 |
HI Andy,
At a constant cutoff the torque on a steam engine is different than an IC engine. The torque climbs to about 3600 then drops one to one hp at 5250.
On a short cutoff the torque is constant where it is hp equal at 5250. Long cutoff low preasure engines have tremendous valve velosities and a preasure drop on the piston at the long cutoffs. Jay carters engine ran at 5000 Rpm at 2000 psi and at 40cu in was about 90 hp. An exellent performance for the day. He also belives in high RPM. All engines drop to hp and torque equal at 5250.
Harry
At a constant cutoff the torque on a steam engine is different than an IC engine. The torque climbs to about 3600 then drops one to one hp at 5250.
On a short cutoff the torque is constant where it is hp equal at 5250. Long cutoff low preasure engines have tremendous valve velosities and a preasure drop on the piston at the long cutoffs. Jay carters engine ran at 5000 Rpm at 2000 psi and at 40cu in was about 90 hp. An exellent performance for the day. He also belives in high RPM. All engines drop to hp and torque equal at 5250.
Harry
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Re: Williams ws. Rankin September 29, 2006 11:22AM |
Registered: 19 years ago Posts: 1,038 |
Hi Harry
HP and torque are not equal at 5250 but they are close. The reasion is that by defination 1 HP = 33000 ft-lb/min and 1 lb-ft torque = 2 pi ft-lb work per revoloution.
33000/(2*pi)=5252.113122032546080373164191293
HP and torque are almost equal at 5252.113122032546080373164191293 RPM
So at 5252.115 RPM torque and HP are very close to equal. That is independent of whatever is producing that torque.
The one modern engine we have data for is the S.E.S. engine. It was a short cutoff high expansion engine. I don't remember the details at the moment. Have the Paper at home. But it's HP peeked around 1650 RPM. We also have the pressure ahead of the inlet valve. And it can be seen how the pressure reacts to the valve opening and closing. You can see the pressure builds when the valve closes and as the flow reflects back the pressure drops and builds as the standing wave bounces back and forth. Tuning the port length leading to the valve changes that frequency and amptitude.
The one difference between the engines I have data on and yours or Jays that might make the differance is pressure. All the engines I have torque data on are using 1000 PSIA or less steam pressure. Except maybe the Doble data might for be 1400 PSIA. But I think Jim or someone told me it wasn't.
Jay told me he ran as high as 2800 PSI.
"At a constant cutoff the torque on a steam engine is different than an IC engine." Well yes Thats a given.
"The torque climbs to about 3600 then drops one to one hp at 5250." Are you talking about your engine here?
Every bit of data I have shows max torque at close to 0 RPM gradually droping in a stright line until at about 900 RPM where it has droped about 5% or less. From 900 to about 1100 it's a curve droping maybe 5 to 10 percent. from there it draps rapidly. By 1800 RPM the torque is about 15% of the (0 RPM) max. By 2500 RPM the torque has droped to around 2 or 3 percent of the (0 RPM) max. The S.E.S. tuned port changed the torque curve decreasing the slop at higher RPM. At 1800 RPM you still had about 30 tp 40 percent of max torque. Peek HP except for the tuned port S.E.S. engine is between 1200 and 1300 RPM. The tuned port moved peek HP to about 1650 RPM.
I don't have hard data. But experanced steamers have told me that small engines run higher RPMs.
TYpical torque curve data look something like below.
*********************
**************************
*****************************
*******************************
*********************************
***********************************
*************************************
****************************************
*******************************************
***********************************************
**********************************************************
0---------------------900---------------------1800
The data I have are for the Doble, White and S.E.S. engines and they all look simular to the above. The tuned port S.E.S. just about halves the decent rate above 900 RPM. At 2800 RPM the torque is about what the untuned engines are at 1800.
You havn't shown me anything that proves different. I am from MO. Show me. Like to see dyno torque or HP chart for your engine.
If you have solved the problem then how. Is it high pressure, or small cylanders, or relation of steam density to cylnder displacement. Is cutoff in the high RPM formula. What gets high RPM?
Andy
Edited 1 time(s). Last edit at 09/29/2006 03:17PM by Andy.
HP and torque are not equal at 5250 but they are close. The reasion is that by defination 1 HP = 33000 ft-lb/min and 1 lb-ft torque = 2 pi ft-lb work per revoloution.
33000/(2*pi)=5252.113122032546080373164191293
HP and torque are almost equal at 5252.113122032546080373164191293 RPM
So at 5252.115 RPM torque and HP are very close to equal. That is independent of whatever is producing that torque.
The one modern engine we have data for is the S.E.S. engine. It was a short cutoff high expansion engine. I don't remember the details at the moment. Have the Paper at home. But it's HP peeked around 1650 RPM. We also have the pressure ahead of the inlet valve. And it can be seen how the pressure reacts to the valve opening and closing. You can see the pressure builds when the valve closes and as the flow reflects back the pressure drops and builds as the standing wave bounces back and forth. Tuning the port length leading to the valve changes that frequency and amptitude.
The one difference between the engines I have data on and yours or Jays that might make the differance is pressure. All the engines I have torque data on are using 1000 PSIA or less steam pressure. Except maybe the Doble data might for be 1400 PSIA. But I think Jim or someone told me it wasn't.
Jay told me he ran as high as 2800 PSI.
"At a constant cutoff the torque on a steam engine is different than an IC engine." Well yes Thats a given.
"The torque climbs to about 3600 then drops one to one hp at 5250." Are you talking about your engine here?
Every bit of data I have shows max torque at close to 0 RPM gradually droping in a stright line until at about 900 RPM where it has droped about 5% or less. From 900 to about 1100 it's a curve droping maybe 5 to 10 percent. from there it draps rapidly. By 1800 RPM the torque is about 15% of the (0 RPM) max. By 2500 RPM the torque has droped to around 2 or 3 percent of the (0 RPM) max. The S.E.S. tuned port changed the torque curve decreasing the slop at higher RPM. At 1800 RPM you still had about 30 tp 40 percent of max torque. Peek HP except for the tuned port S.E.S. engine is between 1200 and 1300 RPM. The tuned port moved peek HP to about 1650 RPM.
I don't have hard data. But experanced steamers have told me that small engines run higher RPMs.
TYpical torque curve data look something like below.
*********************
**************************
*****************************
*******************************
*********************************
***********************************
*************************************
****************************************
*******************************************
***********************************************
**********************************************************
0---------------------900---------------------1800
The data I have are for the Doble, White and S.E.S. engines and they all look simular to the above. The tuned port S.E.S. just about halves the decent rate above 900 RPM. At 2800 RPM the torque is about what the untuned engines are at 1800.
You havn't shown me anything that proves different. I am from MO. Show me. Like to see dyno torque or HP chart for your engine.
If you have solved the problem then how. Is it high pressure, or small cylanders, or relation of steam density to cylnder displacement. Is cutoff in the high RPM formula. What gets high RPM?
Andy
Edited 1 time(s). Last edit at 09/29/2006 03:17PM by Andy.
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Re: Williams ws. Rankin September 29, 2006 03:33PM |
Registered: 19 years ago Posts: 1,441 |
Andy,
You are seeing something that I have been screeching about for years.
1) Poor port design with too much turbulent flow and restriction.
2) The inlet valve is too small and/or is restricted.
White valves were too small for the displacement and long ports, and the Doble E had way too long passages, what with that Woolf compound design.
Any engine with Stanley type porting is going to be bad, they have to be straight in. These engines were choking themselves.
Same goes for unaflow engines, the exhaust ports have to be big. Look at a GM Detroit Diesel sleeve and see what good ports look like.
You have to take a good look at a modern IC race engine to see what is good porting.
There hasn't been a really good steam engine designed yet, except for the Skinner.
JC
You are seeing something that I have been screeching about for years.
1) Poor port design with too much turbulent flow and restriction.
2) The inlet valve is too small and/or is restricted.
White valves were too small for the displacement and long ports, and the Doble E had way too long passages, what with that Woolf compound design.
Any engine with Stanley type porting is going to be bad, they have to be straight in. These engines were choking themselves.
Same goes for unaflow engines, the exhaust ports have to be big. Look at a GM Detroit Diesel sleeve and see what good ports look like.
You have to take a good look at a modern IC race engine to see what is good porting.
There hasn't been a really good steam engine designed yet, except for the Skinner.
JC
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Re: Williams ws. Rankin September 29, 2006 05:00PM |
Registered: 19 years ago Posts: 835 |
Hi guys.
Wish you had come to last SACA meet and seen things first hand. as again you are talking about long cutoff engines. With an oliptical crank throw there is plenty of time to fill the cylinder and let it do its work instead of equilizing port preasure with cylinder down stroke. Last year the fellow from SES was there. his comment was " you are on the right track and far ahead of where we were". That is the difference in a modern engine. the valve timing and duration change at about 1000rpm to a short cutoff and advance timing with an a decrease in clearance volume.
Harry
Wish you had come to last SACA meet and seen things first hand. as again you are talking about long cutoff engines. With an oliptical crank throw there is plenty of time to fill the cylinder and let it do its work instead of equilizing port preasure with cylinder down stroke. Last year the fellow from SES was there. his comment was " you are on the right track and far ahead of where we were". That is the difference in a modern engine. the valve timing and duration change at about 1000rpm to a short cutoff and advance timing with an a decrease in clearance volume.
Harry
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Re: Williams ws. Rankin September 29, 2006 08:08PM |
Registered: 19 years ago Posts: 318 |
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Re: Williams ws. Rankin September 29, 2006 08:46PM |
Registered: 19 years ago Posts: 318 |
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Re: Williams ws. Rankin October 01, 2006 08:20AM |
Registered: 19 years ago Posts: 215 |
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Re: Williams ws. Rankin October 01, 2006 11:08AM |
Registered: 19 years ago Posts: 1,441 |
George,
Indeed the Series E ports were long; but they were also quite large in total area.
Same for the size of the piston valve and it's ports.
What always bothered me was the actual clearance volume in this engine, especially for the LP cylinder. In none of the Doble engineering books is there any mention of the actual clearance volumes.
Remember when Fred Marriott complained about the "Buzzing" or vibration in the original Rocket engine? Then, according to the stories, Stanley altered the cylinder block for 1907.
I often wondered if they didn't enlarge the ports and the exhaust cavity in the valve for this one engine. Then they broke it, so the actual run was done with a standard 30 hp engine. That one probably was indeed choking itself.
JC
Indeed the Series E ports were long; but they were also quite large in total area.
Same for the size of the piston valve and it's ports.
What always bothered me was the actual clearance volume in this engine, especially for the LP cylinder. In none of the Doble engineering books is there any mention of the actual clearance volumes.
Remember when Fred Marriott complained about the "Buzzing" or vibration in the original Rocket engine? Then, according to the stories, Stanley altered the cylinder block for 1907.
I often wondered if they didn't enlarge the ports and the exhaust cavity in the valve for this one engine. Then they broke it, so the actual run was done with a standard 30 hp engine. That one probably was indeed choking itself.
JC
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Re: Williams ws. Rankin October 01, 2006 11:47AM |
Registered: 19 years ago Posts: 171 |
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Re: Williams ws. Rankin October 01, 2006 02:11PM |
Registered: 19 years ago Posts: 1,268 |
Hello Jim & George,
Has anyone done a takeoff on the exact size of both inlet and exhaust ports compared to the volume of steam going through them per stroke for the better running engines? This would have to be done as effective port area related to volume not mass to get usable results.
For instance here I THINK my inlet design, transfer port and exhaust port design are maxed out but don't have much hard comparative data.
Jim you were right in that the second stage needs an inlet valve. Mine is a recieverless unaflow compound with full recompression on both stages, so I had thought since the whole thing expands as a single cylinder that a valve would not be needed. It is not needed to control cutoff for the second stage but to control residual pressure for the first. Ie transfer port opening remains fixed to TDC of the second stage but the closing has to be variable. I suppose that is second stage cutoff come to think of it.
Anyway thanks for mentioning the Burt-McCullom sleeve valve. I am working on a take off from it that will work for this transfer port valve. One advantage of this type of sleeve valve is that there is always some movement between the cylinder and rings, keeping the rings above boundary and into hydrodynamic lubrication. I am a little hesitant to use the actuating mechanism they did with the spherical ball bushing and sliding pin and am working on some better way. Other ways have floated by but nothing really better yet. In any case it has to occult the ports at about a 45 degree angle to vertical to get the fastest combination of rotational and reciprocating ingrediants to snap the port closed as quickly as possible.
My Best to everyone ------------ Bill G.
Edited 1 time(s). Last edit at 10/01/2006 02:12PM by Bill Gatlin.
Has anyone done a takeoff on the exact size of both inlet and exhaust ports compared to the volume of steam going through them per stroke for the better running engines? This would have to be done as effective port area related to volume not mass to get usable results.
For instance here I THINK my inlet design, transfer port and exhaust port design are maxed out but don't have much hard comparative data.
Jim you were right in that the second stage needs an inlet valve. Mine is a recieverless unaflow compound with full recompression on both stages, so I had thought since the whole thing expands as a single cylinder that a valve would not be needed. It is not needed to control cutoff for the second stage but to control residual pressure for the first. Ie transfer port opening remains fixed to TDC of the second stage but the closing has to be variable. I suppose that is second stage cutoff come to think of it.
Anyway thanks for mentioning the Burt-McCullom sleeve valve. I am working on a take off from it that will work for this transfer port valve. One advantage of this type of sleeve valve is that there is always some movement between the cylinder and rings, keeping the rings above boundary and into hydrodynamic lubrication. I am a little hesitant to use the actuating mechanism they did with the spherical ball bushing and sliding pin and am working on some better way. Other ways have floated by but nothing really better yet. In any case it has to occult the ports at about a 45 degree angle to vertical to get the fastest combination of rotational and reciprocating ingrediants to snap the port closed as quickly as possible.
My Best to everyone ------------ Bill G.
Edited 1 time(s). Last edit at 10/01/2006 02:12PM by Bill Gatlin.
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Re: Williams ws. Rankin October 01, 2006 04:10PM |
Registered: 19 years ago Posts: 1,441 |
Bill,
That ball joint business of driving the single sleeve was chosen by all who used it in WW-2, just because it was so simple and also reliable in aircraft service.
There is the ball in it's holder on the bottom of the sleeve, free to revolve. Then the little crank with the overhung crank pin engages the hole in the ball. There is relative movement between the pin and the ball.
Correct on saying that the sleeve is in constant motion. That is why the Burt-McCollum single sleeve was infinitely superior to the Knight double sleeve valve.
Sir Harry Ricardo's book, "The High Speed Internal Combustion Engine", 1953-4-5 editions is an absolute gold mine of information on this valve and using it. He pushed it to the absolute limits.
The minimum clearance volume possible, and the largest ports with maximum flow rate over any two or four valve ideas, total control over the head end shape.
JC
Edited 1 time(s). Last edit at 10/02/2006 06:38AM by James D. Crank.
That ball joint business of driving the single sleeve was chosen by all who used it in WW-2, just because it was so simple and also reliable in aircraft service.
There is the ball in it's holder on the bottom of the sleeve, free to revolve. Then the little crank with the overhung crank pin engages the hole in the ball. There is relative movement between the pin and the ball.
Correct on saying that the sleeve is in constant motion. That is why the Burt-McCollum single sleeve was infinitely superior to the Knight double sleeve valve.
Sir Harry Ricardo's book, "The High Speed Internal Combustion Engine", 1953-4-5 editions is an absolute gold mine of information on this valve and using it. He pushed it to the absolute limits.
The minimum clearance volume possible, and the largest ports with maximum flow rate over any two or four valve ideas, total control over the head end shape.
JC
Edited 1 time(s). Last edit at 10/02/2006 06:38AM by James D. Crank.
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Re: Williams ws. Rankin October 01, 2006 08:18PM |
Registered: 19 years ago Posts: 1,268 |
Thanks Jim,
I will try to get a copy of the book. Looking at the geometry of the drive mechanism the ball would be moving toward and away from the crank pin as the sleeve turned, that is why I thought that it must slide on the pin. Evidently then the ball clearances can take up the in and out movement. If it is good enough for aircraft service than it should work for a long life steam engine.
I am hoping this is the last major design hurtle (it won't be) mechanically. The problem is that things like this modify the layout and dimentions a bit. Soon I should have something close enough to working drawings to start a thourough thermo analysis of the thing. I believe I have found some software that I can afford that will do the flow, turbulence and heat transfer dynamics on the paper engine. All in living color.
Getting exiting ---------- Bill G.
I will try to get a copy of the book. Looking at the geometry of the drive mechanism the ball would be moving toward and away from the crank pin as the sleeve turned, that is why I thought that it must slide on the pin. Evidently then the ball clearances can take up the in and out movement. If it is good enough for aircraft service than it should work for a long life steam engine.
I am hoping this is the last major design hurtle (it won't be) mechanically. The problem is that things like this modify the layout and dimentions a bit. Soon I should have something close enough to working drawings to start a thourough thermo analysis of the thing. I believe I have found some software that I can afford that will do the flow, turbulence and heat transfer dynamics on the paper engine. All in living color.
Getting exiting ---------- Bill G.
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Re: Williams ws. Rankin October 01, 2006 10:04PM |
Registered: 19 years ago Posts: 1,441 |
Bill,
The ball oscillates in it's clamp housing at the bottom of the sleeve. There is relative movement between the ball and the overhung crank pin on the little crankshaft; but it is certainly minor.
There were dozens of ways to drive the sleeve that were tried out in the 1920's for aircraft engines, mainly by Bristol and Napier. Lots of them work just fine; but it seems as if the ball on the stud idea was the most reliable and easiest to make.
None of the many reports and flightline service information mention any problem with the drive end of things.
JC
Edited 1 time(s). Last edit at 10/02/2006 06:42AM by James D. Crank.
The ball oscillates in it's clamp housing at the bottom of the sleeve. There is relative movement between the ball and the overhung crank pin on the little crankshaft; but it is certainly minor.
There were dozens of ways to drive the sleeve that were tried out in the 1920's for aircraft engines, mainly by Bristol and Napier. Lots of them work just fine; but it seems as if the ball on the stud idea was the most reliable and easiest to make.
None of the many reports and flightline service information mention any problem with the drive end of things.
JC
Edited 1 time(s). Last edit at 10/02/2006 06:42AM by James D. Crank.
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Re: Williams ws. Rankin October 02, 2006 12:48AM |
Registered: 19 years ago Posts: 1,268 |
Jim,
Is the ball or any of the parts made of brass or is everything hardened steel? Also wondering what size the ball is.
The movement between the ball and pin would be minor but of the order of roughly 1/4" for a 5" dia cylinder depending on the throw of the crank mechanism. Because of the curvature of the valve cylinder it seems as if something would have to move in and out somewhere, or the crank arm is hinged to the shaft.
Appreciate the info -------- Bill G.
Is the ball or any of the parts made of brass or is everything hardened steel? Also wondering what size the ball is.
The movement between the ball and pin would be minor but of the order of roughly 1/4" for a 5" dia cylinder depending on the throw of the crank mechanism. Because of the curvature of the valve cylinder it seems as if something would have to move in and out somewhere, or the crank arm is hinged to the shaft.
Appreciate the info -------- Bill G.
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Re: Williams ws. Rankin October 02, 2006 06:43AM |
Registered: 19 years ago Posts: 1,441 |
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Re: Williams ws. Rankin October 02, 2006 01:20PM |
Registered: 19 years ago Posts: 1,038 |
Hi Howard
Tnks. The end of expansion temperature is dependent on the initial temperature, pressure and expansion raio. The exhaust temperature I am not sure how to figure as there is some debate on what process for the pressure drop to exhaust pressure and the result comes out quite different. I did a quick analysis for a 5 x expansion from 850F figured the exhaust temperature is a constant enthalpy process and to obay conservation of energy law. The result is quite different.
I am attaching the analysis as a PDF document. I also did it for 27 to 1 expansion ratio. I did them over a pressure range. 400 to 4500 PSIA for the 5X expansion and 850 to 4500 PSIA for the 27X expansion. 850 PSIA avoids over expansion cases. As you can see the higher expansion ratio requires more super heat to avoid to wet of an exhaust steam.
My three stage compound design avoids wet exhaust by design. As each stage uses a bit more steam then the previous a bit of high temp steam is mixed in to make up the additional amount used. That adds heat to the steam used by successive stages.
Andy
Tnks. The end of expansion temperature is dependent on the initial temperature, pressure and expansion raio. The exhaust temperature I am not sure how to figure as there is some debate on what process for the pressure drop to exhaust pressure and the result comes out quite different. I did a quick analysis for a 5 x expansion from 850F figured the exhaust temperature is a constant enthalpy process and to obay conservation of energy law. The result is quite different.
I am attaching the analysis as a PDF document. I also did it for 27 to 1 expansion ratio. I did them over a pressure range. 400 to 4500 PSIA for the 5X expansion and 850 to 4500 PSIA for the 27X expansion. 850 PSIA avoids over expansion cases. As you can see the higher expansion ratio requires more super heat to avoid to wet of an exhaust steam.
My three stage compound design avoids wet exhaust by design. As each stage uses a bit more steam then the previous a bit of high temp steam is mixed in to make up the additional amount used. That adds heat to the steam used by successive stages.
Andy
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Re: Williams ws. Rankin October 03, 2006 10:18AM |
Registered: 19 years ago Posts: 835 |
Hi Andy,
Sorry a typo on the 5250 should have been 5252. This the standard for computing the torque / RPM to hp from a dyno. A dyno will read the torque and RPM.
It is RPM / 5252 x torque = hp of course you know this
5000rpm / 5252= .952 x400torque = 380.8hp
The bump valve can operate at the higher rpm range so we rated our two cyl at 4400rpm where as the 6cyl is rated at 3600 because of the valve train. Still working on a lighter system to increase the rpm. Bill can relate to this.
Harry
Sorry a typo on the 5250 should have been 5252. This the standard for computing the torque / RPM to hp from a dyno. A dyno will read the torque and RPM.
It is RPM / 5252 x torque = hp of course you know this
5000rpm / 5252= .952 x400torque = 380.8hp
The bump valve can operate at the higher rpm range so we rated our two cyl at 4400rpm where as the 6cyl is rated at 3600 because of the valve train. Still working on a lighter system to increase the rpm. Bill can relate to this.
Harry
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