Williams ws. Rankin
Posted by Howard Langdon
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Re: Williams ws. Rankin December 03, 2005 06:43PM |
Registered: 18 years ago Posts: 1,249 |
Hello Harry,
Thanks for the tip, I'll write them for more info.
I agree on the neccessity of High compression. Without it the higher efficiencys are just too hard to get. That's why both stages of the compound are recompressing to inlet conditions.
I still see the reasons for the extra expansion that the compound offers as also necessary for high efficiency. Jerry's papers are quite illuminating.
The cam design is coming along nicely. Actually I am glad I got sidetracked onto it. What I wasn't realising was how much the total valve lift is compromised around short cut off, and the degree duration of any valve movement in general. How long to open, how long to close, and the low amount of lift when these two events come close together around short cut off. All parameters that the engine will have to be designed around.
One of the ideas, I think, that people may have mistakenly gotten is through making a comparison of steam engine valve dynamics with those of the gasoline or diesel engines. In these engines the inlet or exhaust valve openings and closings are covering a very wide number of crankshaft degrees. From just before TDC to just after BDC with the intake for instance. A steam engine is trying for from say sixty to zero degrees. A minimum of three to many times faster. Some serious dynamics.
Square lobes don't work, not even with square rollers, what was Andy saying about timeing there?
Hopeing the best for everybody ---------Bill G.
Thanks for the tip, I'll write them for more info.
I agree on the neccessity of High compression. Without it the higher efficiencys are just too hard to get. That's why both stages of the compound are recompressing to inlet conditions.
I still see the reasons for the extra expansion that the compound offers as also necessary for high efficiency. Jerry's papers are quite illuminating.
The cam design is coming along nicely. Actually I am glad I got sidetracked onto it. What I wasn't realising was how much the total valve lift is compromised around short cut off, and the degree duration of any valve movement in general. How long to open, how long to close, and the low amount of lift when these two events come close together around short cut off. All parameters that the engine will have to be designed around.
One of the ideas, I think, that people may have mistakenly gotten is through making a comparison of steam engine valve dynamics with those of the gasoline or diesel engines. In these engines the inlet or exhaust valve openings and closings are covering a very wide number of crankshaft degrees. From just before TDC to just after BDC with the intake for instance. A steam engine is trying for from say sixty to zero degrees. A minimum of three to many times faster. Some serious dynamics.
Square lobes don't work, not even with square rollers, what was Andy saying about timeing there?
Hopeing the best for everybody ---------Bill G.
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Re: Williams ws. Rankin December 03, 2005 07:12PM |
Registered: 18 years ago Posts: 171 |
One simple thing to note,,, A gas engine intake is propeld by 15 pounds ,, but the steam is propeld by,,er,,ahh say 600 + pounds,,,so naturaly the lift need be smaller and of fewer degrees And to get it out again,,more lift and time is needed,,The great advantage of uniflow and compound is by keeping the hot end hot,,and the cold end cold,,,and the steam going one way,,,the rest are details,,, Expansion without losing heat to a cold wall is the problem ,,,Hope I didn't overstate it,,,Cheers Ben
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Re: Williams ws. Rankin December 04, 2005 12:54AM |
Registered: 18 years ago Posts: 1,249 |
Thanks Ben,
I had hoped to get a 1/2 inch valve lift. What I am finding is that at my 30% maximum cutoff is that that lift, if reachable at all, is a peak in the middle. As cutoff reduces the peak drops.
Because of some advance and not much piston movement around TDC the opening lift doesn't seem too critical but a slow closing at the end of cutoff when the piston is moving more rapidly will have a more serious effect with wire drawing.
I'll know more when I am finished with the geometry and can put some real mass and inertia figures together to see what lift rates are reasonable at what RPMs.
The whole affair is a take off on the Skinner duel cam system. I had a cam that could change configuration well up to the point where the up slope and the down slope started to overlap. That left too sharp a discontinuity for the roller to traverse smoothly. Turning my hydraulic cam system idea into a mechanical system ended up with two cams working much like the Skinner.
It's a lot of fun work.
Be well ------------- Bill G.
I had hoped to get a 1/2 inch valve lift. What I am finding is that at my 30% maximum cutoff is that that lift, if reachable at all, is a peak in the middle. As cutoff reduces the peak drops.
Because of some advance and not much piston movement around TDC the opening lift doesn't seem too critical but a slow closing at the end of cutoff when the piston is moving more rapidly will have a more serious effect with wire drawing.
I'll know more when I am finished with the geometry and can put some real mass and inertia figures together to see what lift rates are reasonable at what RPMs.
The whole affair is a take off on the Skinner duel cam system. I had a cam that could change configuration well up to the point where the up slope and the down slope started to overlap. That left too sharp a discontinuity for the roller to traverse smoothly. Turning my hydraulic cam system idea into a mechanical system ended up with two cams working much like the Skinner.
It's a lot of fun work.
Be well ------------- Bill G.
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Re: Williams ws. Rankin December 05, 2005 01:52PM |
Registered: 18 years ago Posts: 1,038 |
Hi Ben
I am aware of the uniflow hype. But I am not so sure the theory is correct. Read Jerry's papers on the Williams engines. There is a whole lot to be gained from high compression. When using accepted IFC-67 steam properties in analyzing high compression and Jerry's principals you wind up with efficiencies very close to real engines. Compression reclames more then is lost to clearance acording to Jerry's papers.
Using Jerry's constant enthalpy pressure drop process at the end of expansion to exhaust pressure and calculating a compression cycle with 1000 PSIA at 650F
Inlet 1000PSIA 650F H=1290.1BTU
Cutoff 1000PSIA 756.4F H=1362.7BTU
Expansion 19.5PSIA 226.3F H=1026.3BTU
Residual 14.696PSIA 212.0F H=1026.3
compression 1000.9PSIA 809.2F H=1395.1BTU
The steam at cutoff is a mix of 30.9% inlet and 69.1% residual steam.
But to achieve the above the cutoff is 1% and clearance is 2.75% Exhaust is 10% of stroke.
The problem is that cutoff is just to dam short. About 10 degrees of crank rotation. At 600 RPM, 10 revoloutions per second. One revoloution takes 1/10 of a second. The inlet valve must open and close in 10/360 of 0.10 seconds.
At 600 RPM you have 0.00277777 seconds for the inlet valve to open and close. With a little advance on opening maybe around 0.003 seconds. Anyways it is to short time wise. Thus the reasion fo going to a compound engine. We can get the same expansion in a three stage compound with around 30% cutoff. Though more of the compounds expansion is non-work producing. The interstage pressure drops, end of expansion to exhaust pressure, produce no work.
I am aware of the uniflow hype. But I am not so sure the theory is correct. Read Jerry's papers on the Williams engines. There is a whole lot to be gained from high compression. When using accepted IFC-67 steam properties in analyzing high compression and Jerry's principals you wind up with efficiencies very close to real engines. Compression reclames more then is lost to clearance acording to Jerry's papers.
Using Jerry's constant enthalpy pressure drop process at the end of expansion to exhaust pressure and calculating a compression cycle with 1000 PSIA at 650F
Inlet 1000PSIA 650F H=1290.1BTU
Cutoff 1000PSIA 756.4F H=1362.7BTU
Expansion 19.5PSIA 226.3F H=1026.3BTU
Residual 14.696PSIA 212.0F H=1026.3
compression 1000.9PSIA 809.2F H=1395.1BTU
The steam at cutoff is a mix of 30.9% inlet and 69.1% residual steam.
But to achieve the above the cutoff is 1% and clearance is 2.75% Exhaust is 10% of stroke.
The problem is that cutoff is just to dam short. About 10 degrees of crank rotation. At 600 RPM, 10 revoloutions per second. One revoloution takes 1/10 of a second. The inlet valve must open and close in 10/360 of 0.10 seconds.
At 600 RPM you have 0.00277777 seconds for the inlet valve to open and close. With a little advance on opening maybe around 0.003 seconds. Anyways it is to short time wise. Thus the reasion fo going to a compound engine. We can get the same expansion in a three stage compound with around 30% cutoff. Though more of the compounds expansion is non-work producing. The interstage pressure drops, end of expansion to exhaust pressure, produce no work.
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Re: Williams ws. Rankin December 05, 2005 04:24PM |
Registered: 18 years ago Posts: 835 |
Hi Andy,
Would you not have the same effect using a long stroke to bore ratio ie 2" bore x 6" stroke as this will allow for more expansions and maintaining the same piston speed will decrease the rpm for the same power also longer valve timing and less porting. of course operating at 650f there will be a low expansion ratio. With high preasure and high temp where modern engines are running where is the gain in relation to the complication and conductive losses?
Harry
Would you not have the same effect using a long stroke to bore ratio ie 2" bore x 6" stroke as this will allow for more expansions and maintaining the same piston speed will decrease the rpm for the same power also longer valve timing and less porting. of course operating at 650f there will be a low expansion ratio. With high preasure and high temp where modern engines are running where is the gain in relation to the complication and conductive losses?
Harry
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Re: Williams ws. Rankin December 05, 2005 04:50PM |
Registered: 18 years ago Posts: 171 |
Hi Andy,,,I'LL re read and see if I can say something useful ,,,long day today,,,Bill,,,Re valves and lift,,which is so important to lo pressure otto cycle engines,,,Back some years ago,,someone at MIT,,wrote a senior? thesis titled[?] ''Flow through popppet valves at various lifts and pressures'' I cant recall if it was before the war or after,,,this paper is much referd to in tech journals,,and if you go to MIT,,its even more fun to see who has signed in to have access to it for a hour or 3 ,,,in a chain link cage wey deep in the middle of some building off Mass ave,,,Anyhow ,, My take is ya want a lift of 25 percent of the valve diameter,,,now this as it applies to otto eng' and their related cams etc,,XXXX Now on poppet valves,,I have found Chrysler,hemi stuff with the higher spring loads for the heavier rocker gear,,,the valve material is a better grade,,and will stand up better,,This experiance gained using 2, 3/32 dia intake valves ,,thined a lot and a 1/4 inch hole drilled down the stem,,,ya reallly,,haha,,,You could barely see the plug hole in the head,,,It was a fun day to have friends guess which valve was heavier,,,and the largest one was the lightest,,,oh yes one of em was titanium,,came special order for 35 bucks,in a little brown bag special delivery right from the factory,,thompson I think,,,fall of '69,,,Now that MIT paper was not intended to be more than reserch on the problem at hand,, so they were not concernd w/ valve gear or duration,,,but rather generate a formula that would apply to a wide virety of conditions, constant conditions,,,,,hmmmm,,,,Tipi'n is NOT my forte,,grr,,,Cheers Ben p/s we did not use the titanium valve,,Poor heat flow for ex,,and the steel in would stay cooler and be lighter on the intake,,haha
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Re: Williams ws. Rankin December 05, 2005 04:58PM |
Registered: 18 years ago Posts: 171 |
Oh , I forgot,,Make your valve heads w/no curves,,sacrifice flow for straight lines,,,this way if the valve BENDS,,you can see it by laying a straight edge across it,,after you get past the stage of keeping it together,,then and only then go exotic,My valve cracked a truncated pie shaped piece,,right through,,,John finnished 3rd in the 50miler at Montreal that day,I think,,,said she was flat the last 3 laps,,Bike 1950 Vincent gray flash,,race 1970,,Ben
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Re: Williams ws. Rankin December 05, 2005 05:49PM |
Registered: 18 years ago Posts: 1,038 |
Hi Herry
The time duration of the event is what I am looking at. Did a quick calk changing bore stroke relation. The cutoff on degrease decreased. Left conecting rod length same. Assumming enough flywheel initia the angular velocity is relative constant. Changing bore and stroke would have little change on timming. No change unless the conecting rod length to stroke length changes. The ratio of the conecting rod length and stroke length effect timming relation of degrees to percent of stroke.
But there is little that can be done there to get significant more time.
We are talking about a few tenths of a degree. 1% cutoff varies from about 9.5 to 11.5 depending rod length to stroke ratio.
The gain is actually obtaining thoes very high expansions. It's no sure thing in any case. But there losses either way you go. The conduction losses are there in both cases. Conduction/convection losses are dependent on surface area. And it simple to look at that. You have x amount of steam going through the engine producing power. Looking at a compound engine vs a simple engine. Say we can get the same expansion from both. In the end we have some volume of steam. The volume is constrained by surfaces. The ending volumes will be simular so the two engine would have simular over all surface areas and thus simular conduction losses. But wait. What of the internal heat transfer. In the simple engine we have huge over temperature differance in a single expansion driving a higher rate of internal heat transfer. While staged we have reduced that temperature differances into steps. The temperature change with a stage is significantly less and thus less heat conduction losses. We will need revievers between stages unless we have steaple timming. Thoes devices should have little conduction losses as they are at constant temperature and can be well insolated.
We have history to go by. We have compound engines and uniflow engines having close efficiencies. The uniflow generally beats the compound. But what happens when we have a high compression compound. That brings a stage up to the uniflows efficiency level by eliminating clearance losses. We then just have the interstage isenthalip pressure drops losses. So the question boiles down to the interstage losses vs the wire drawing losses of the very short cutoff of a simple engine.
I am not so sure that a simple engine can ever get valve operation down to the cutoff needed for controling power range I described in other posts.
Andy
The time duration of the event is what I am looking at. Did a quick calk changing bore stroke relation. The cutoff on degrease decreased. Left conecting rod length same. Assumming enough flywheel initia the angular velocity is relative constant. Changing bore and stroke would have little change on timming. No change unless the conecting rod length to stroke length changes. The ratio of the conecting rod length and stroke length effect timming relation of degrees to percent of stroke.
But there is little that can be done there to get significant more time.
We are talking about a few tenths of a degree. 1% cutoff varies from about 9.5 to 11.5 depending rod length to stroke ratio.
The gain is actually obtaining thoes very high expansions. It's no sure thing in any case. But there losses either way you go. The conduction losses are there in both cases. Conduction/convection losses are dependent on surface area. And it simple to look at that. You have x amount of steam going through the engine producing power. Looking at a compound engine vs a simple engine. Say we can get the same expansion from both. In the end we have some volume of steam. The volume is constrained by surfaces. The ending volumes will be simular so the two engine would have simular over all surface areas and thus simular conduction losses. But wait. What of the internal heat transfer. In the simple engine we have huge over temperature differance in a single expansion driving a higher rate of internal heat transfer. While staged we have reduced that temperature differances into steps. The temperature change with a stage is significantly less and thus less heat conduction losses. We will need revievers between stages unless we have steaple timming. Thoes devices should have little conduction losses as they are at constant temperature and can be well insolated.
We have history to go by. We have compound engines and uniflow engines having close efficiencies. The uniflow generally beats the compound. But what happens when we have a high compression compound. That brings a stage up to the uniflows efficiency level by eliminating clearance losses. We then just have the interstage isenthalip pressure drops losses. So the question boiles down to the interstage losses vs the wire drawing losses of the very short cutoff of a simple engine.
I am not so sure that a simple engine can ever get valve operation down to the cutoff needed for controling power range I described in other posts.
Andy
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Re: Williams ws. Rankin December 05, 2005 06:13PM |
Registered: 18 years ago Posts: 1,038 |
Hi Ben
Hoped up motocycles in my past also, But we are looking at steam engins cutoff of very low percentage of stroke in a simple engine to get high expansions. In the example above a came up with a 3ms valve open duration for around a 27:1 expansion ratio. But an IC enging running at the same speed would have a open duration greater then 50ms. 17 times as long as the high expansion steam engine.
600 RPM steam equilivant to 11448 RPM IC inlet valve time duration. So bring the steam engine RPM up and were does that put the time duration. Try to get 3000 RPM out of that steam engine at 27:1 expansion. Like trying to run an IC engine at 57230 RPM.
Herry
At supper critical your steam is a lot more dense. I doubt you are fully expanding down to atmospheric pressure. The thing is that with that high of a pressure you wont be able to fully expand it down atmospheric pressure in a single stage. But on the other hand you can expand a resionable amount and because you are not near the lower pressure limit of expansion at full inlet pressure there is a great deal of room for throttling and you have great potential for high expansion. The valve open duration is less of a problem. So it throttles down. You started so high, that wire drawing is only a small loss. What I am getting at is that using verry high pressure even though you might be able to only expand say 20:1 you get the efficiency of that expansion and have room for throttling.
Andy
Hoped up motocycles in my past also, But we are looking at steam engins cutoff of very low percentage of stroke in a simple engine to get high expansions. In the example above a came up with a 3ms valve open duration for around a 27:1 expansion ratio. But an IC enging running at the same speed would have a open duration greater then 50ms. 17 times as long as the high expansion steam engine.
600 RPM steam equilivant to 11448 RPM IC inlet valve time duration. So bring the steam engine RPM up and were does that put the time duration. Try to get 3000 RPM out of that steam engine at 27:1 expansion. Like trying to run an IC engine at 57230 RPM.
Herry
At supper critical your steam is a lot more dense. I doubt you are fully expanding down to atmospheric pressure. The thing is that with that high of a pressure you wont be able to fully expand it down atmospheric pressure in a single stage. But on the other hand you can expand a resionable amount and because you are not near the lower pressure limit of expansion at full inlet pressure there is a great deal of room for throttling and you have great potential for high expansion. The valve open duration is less of a problem. So it throttles down. You started so high, that wire drawing is only a small loss. What I am getting at is that using verry high pressure even though you might be able to only expand say 20:1 you get the efficiency of that expansion and have room for throttling.
Andy
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Re: Williams ws. Rankin December 05, 2005 09:09PM |
Admin Registered: 18 years ago Posts: 410 |
Gimmee a break Andy.
>>600 RPM steam equilivant to 11448 RPM IC inlet valve time duration. So bring the steam engine RPM up and were does that put the time duration. Try to get 3000 RPM out of that steam engine at 27:1 expansion. Like trying to run an IC engine at 57230 RPM. <<
You are throwing numbers around that are quite meaningless, you provide no context at all.
0, 2 or 4 cycle?
1, We aren't dealing with 15psi absolute steam. Pressure, temp, density, quality?
2, What is the port size?
3, What is the swept volume?
4, Beginning temperature...?
5, Ending temperature...?
6, Pressure differential across the valve for every fractional degree during the duration of valve opening...?
7, Surface area, wall thickness, thermal conductivity.
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Re: Williams ws. Rankin December 05, 2005 10:31PM |
Registered: 18 years ago Posts: 1,249 |
Hello Ben and everyone,
Yeah I have read 30% of the valve diameter for a decent amount of lift in an IC engine. This from one of the camshaft and racing websites I believe.
I am planning on a 7/8 inch valve opening diameter. The obvious differences of course are that the pressures are much higher for the inlet valves at least and that the valves lift into the steamchest to open instead of descend. The valve head is rounded and streamlined for flow. (idea so far)
One thing that is starting to become obvious with the cam timeing is that full opening of the valve will probably be a rare event in the middle of maximum cutoff. (30% in my design) All shorter cutoffs will have to work with much smaller lifts. This can probably be compensated for with higher inlet pressures to some degree at least at higher RPMs.
The cam profiles and linkages can of course be designed to do about anything we might want. Like with the IC engines though it is the various inertial effects and forces that determine the final RPM outcome. Look for instance at the Bump valve. A very light valve that is forced open by the piston at TDC or near but only a few thousandths of an inch, around 20 - 30 as I understand it. I believe this type of valve gets very high RPMs at a quite short cutoff.
When the Williams engine was cruiseing down the highway at short cutoff, what was the timeing events going on then? How much lift and duration at what RPM? Was there any float? As I understand it the Williams low cutoff was around 5% and lift was less than a tenth of an inch, RPMs around three thousand, valve diameter around 5/8 ths, about 5 deg advance. Wheather this all belongs in the same pile I'm not sure.
The thing is that it seems that reasonable or at least workable valve events are to be had down to the 5% cutoff area with good RPMs. Below 5% we are operating so close to TDC that wouldn't we be more interested in just the amount of steam that entered during inlet, without much reguard for the exact timeing of that inlet event? (that's a question)
What I see as probably happening is that under a certain minimum cutoff, say the 5%, all that can be done is to limit the valve lift in the middle of that event, thus controling the mass of entering steam. Would still beat throttling.
Regards to Everyone ------------ Bill G.
Yeah I have read 30% of the valve diameter for a decent amount of lift in an IC engine. This from one of the camshaft and racing websites I believe.
I am planning on a 7/8 inch valve opening diameter. The obvious differences of course are that the pressures are much higher for the inlet valves at least and that the valves lift into the steamchest to open instead of descend. The valve head is rounded and streamlined for flow. (idea so far)
One thing that is starting to become obvious with the cam timeing is that full opening of the valve will probably be a rare event in the middle of maximum cutoff. (30% in my design) All shorter cutoffs will have to work with much smaller lifts. This can probably be compensated for with higher inlet pressures to some degree at least at higher RPMs.
The cam profiles and linkages can of course be designed to do about anything we might want. Like with the IC engines though it is the various inertial effects and forces that determine the final RPM outcome. Look for instance at the Bump valve. A very light valve that is forced open by the piston at TDC or near but only a few thousandths of an inch, around 20 - 30 as I understand it. I believe this type of valve gets very high RPMs at a quite short cutoff.
When the Williams engine was cruiseing down the highway at short cutoff, what was the timeing events going on then? How much lift and duration at what RPM? Was there any float? As I understand it the Williams low cutoff was around 5% and lift was less than a tenth of an inch, RPMs around three thousand, valve diameter around 5/8 ths, about 5 deg advance. Wheather this all belongs in the same pile I'm not sure.
The thing is that it seems that reasonable or at least workable valve events are to be had down to the 5% cutoff area with good RPMs. Below 5% we are operating so close to TDC that wouldn't we be more interested in just the amount of steam that entered during inlet, without much reguard for the exact timeing of that inlet event? (that's a question)
What I see as probably happening is that under a certain minimum cutoff, say the 5%, all that can be done is to limit the valve lift in the middle of that event, thus controling the mass of entering steam. Would still beat throttling.
Regards to Everyone ------------ Bill G.
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Re: Williams ws. Rankin December 06, 2005 10:32AM |
Registered: 18 years ago Posts: 1,441 |
Bill,
We all would certainly like to see verified timing plots for the Williams engine under varying loads and speeds. I doubt that this will ever happen, unless someone is sitting on their note books and data, and will reveal them.
In your deliberations, do remember that one can control either the lift of the valve, or the time it stays open. I have seen both ways of getting cutoff used successfully.
Also take a look at the engine that was designed for the G.M. Pontiac conversion, at least I think that was the engine. They used two poppet valves in series, one to provide the actual event timing and the other one to provide the cutoff point.
What I didn't like about that one was the increase in clearance volume.
One can also take the other approach, use a larger displacement engine and run it slower and thus minimize the valve inertia problem, that is if it really is a problem in the speed range that a steam engine uses for automotive use to begin with.
NASCAR engines run quite well at 9,500 rpm (4750 cam speed which is quite ferocious), at least for a short time, and they do this all day. Let alone the F-1 engines with their pneumatic valve "springs" which now are good for 20K rpm. I doubt a steam engine has to begin to run this fast.
JC
We all would certainly like to see verified timing plots for the Williams engine under varying loads and speeds. I doubt that this will ever happen, unless someone is sitting on their note books and data, and will reveal them.
In your deliberations, do remember that one can control either the lift of the valve, or the time it stays open. I have seen both ways of getting cutoff used successfully.
Also take a look at the engine that was designed for the G.M. Pontiac conversion, at least I think that was the engine. They used two poppet valves in series, one to provide the actual event timing and the other one to provide the cutoff point.
What I didn't like about that one was the increase in clearance volume.
One can also take the other approach, use a larger displacement engine and run it slower and thus minimize the valve inertia problem, that is if it really is a problem in the speed range that a steam engine uses for automotive use to begin with.
NASCAR engines run quite well at 9,500 rpm (4750 cam speed which is quite ferocious), at least for a short time, and they do this all day. Let alone the F-1 engines with their pneumatic valve "springs" which now are good for 20K rpm. I doubt a steam engine has to begin to run this fast.
JC
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Re: Williams ws. Rankin December 06, 2005 11:31AM |
Registered: 18 years ago Posts: 1,038 |
Hi Jim
The S.E.S. engine used series poppet valves. Maybe that is the one you were thinking of.
I have an idea using two cylanderical valves over one port. They alternate one opening and the other closing. Like two slide valves. The covering slide moves off the port opening. The other moves over the port closing it. It's a bit complicated. But seams like it would solve the clearance problem. They would have to be setting right on the head to minimize clearance.
Andy
The S.E.S. engine used series poppet valves. Maybe that is the one you were thinking of.
I have an idea using two cylanderical valves over one port. They alternate one opening and the other closing. Like two slide valves. The covering slide moves off the port opening. The other moves over the port closing it. It's a bit complicated. But seams like it would solve the clearance problem. They would have to be setting right on the head to minimize clearance.
Andy
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Re: Williams ws. Rankin December 06, 2005 02:58PM |
Registered: 18 years ago Posts: 1,249 |
Hi Guys,
One thing to remember is that the cutoff is a percentage of cylinder volume and doesn't relate directly to crankshaft degrees.
Starting at TDC as zero crankshaft degrees, the cylinder volume is: 1-cos$/2
That ends up with my maximum cutoff of 30% for instance at 66.4 degrees of crank rotation from TDC. About 1/2 the duration of an IC engine.
5% cutoff is then 25.8 degrees and 1% is 11.5 degrees.
Say that the IC engine could do its valve lift-dwell-reseat in 120 degrees of crank rotation. I think we might ignore the fact that the IC camshaft is rotating at 1/2 crank speed. The reason is that the valve is shut for approximately half of the time. During that half time another valve event could occur. So our comparison would be at full RPM of the IC engine instead of half, wouldn't it?
So for a cam system designed for full opening in the middle of 66 crankshaft degrees instead of 120 we are looking at about half the RPM of an IC engine without float and the same valve lift. At earlier cutoffs then, the lift is being reduced and the accelerations are the same so no difference in RPM.
In addition, steam engine inlet valves don't have to open as far as IC engine valves do. This, all in all, should get us to a little over 1/2 of the RPM limits of the valve train in an IC engine.
If the valve lift in a IC engine is 0.4 inchs and the valve lift in a steam engine is 0.2 inchs and the valve train acceleration is the same, (the distance covered during acceleration is proportional to the square of the elapsed time) then half of the lift will take 0.707 of the amount of time that full lift would take or 0.707 of the crankshaft degrees. This would put our valve train limit at full cutoff of 30% closer to 1/2 * 1/.707 of that of the IC engine. About 65%
Now the force on the valve train is F=MA, the acceleration is proportional to the velocity squared, which is proportional to RPMs squared. So the same engine has 1/9th of the valve train forces at 3000 RPM that it would at 9000 RPM.
Somehow this all seems to add up to "We Can Do It".
Best to Everyone -------- Bill G.
One thing to remember is that the cutoff is a percentage of cylinder volume and doesn't relate directly to crankshaft degrees.
Starting at TDC as zero crankshaft degrees, the cylinder volume is: 1-cos$/2
That ends up with my maximum cutoff of 30% for instance at 66.4 degrees of crank rotation from TDC. About 1/2 the duration of an IC engine.
5% cutoff is then 25.8 degrees and 1% is 11.5 degrees.
Say that the IC engine could do its valve lift-dwell-reseat in 120 degrees of crank rotation. I think we might ignore the fact that the IC camshaft is rotating at 1/2 crank speed. The reason is that the valve is shut for approximately half of the time. During that half time another valve event could occur. So our comparison would be at full RPM of the IC engine instead of half, wouldn't it?
So for a cam system designed for full opening in the middle of 66 crankshaft degrees instead of 120 we are looking at about half the RPM of an IC engine without float and the same valve lift. At earlier cutoffs then, the lift is being reduced and the accelerations are the same so no difference in RPM.
In addition, steam engine inlet valves don't have to open as far as IC engine valves do. This, all in all, should get us to a little over 1/2 of the RPM limits of the valve train in an IC engine.
If the valve lift in a IC engine is 0.4 inchs and the valve lift in a steam engine is 0.2 inchs and the valve train acceleration is the same, (the distance covered during acceleration is proportional to the square of the elapsed time) then half of the lift will take 0.707 of the amount of time that full lift would take or 0.707 of the crankshaft degrees. This would put our valve train limit at full cutoff of 30% closer to 1/2 * 1/.707 of that of the IC engine. About 65%
Now the force on the valve train is F=MA, the acceleration is proportional to the velocity squared, which is proportional to RPMs squared. So the same engine has 1/9th of the valve train forces at 3000 RPM that it would at 9000 RPM.
Somehow this all seems to add up to "We Can Do It".
Best to Everyone -------- Bill G.
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Re: Williams ws. Rankin December 06, 2005 06:26PM |
Registered: 18 years ago Posts: 1,038 |
Hi Bill
I'm not sure what you are saying on RPM there.
The way I see it is, as a time function of the valve movement. The fact that an IC valve only operates every other revolution is not significant. The rotating cam at half speed of engine shaft is the only differance. The valves lift and close within so many degreas of rotation. Every revoloution or every other revoloution the timming is tied to the events revoloution. There harmonics I suppose effected by the period of the event. I am not sure where you got 120 degrees. An IC valve open before TDC and closes after TDC a full 180+ degrees.
Your calculation of degrees to cutoff is wrong for a normal crank engine. I don't have time to find the formula right now. It is a bit more complex though. My steam cycle calculator has it built in. Gives cutoff in degrees and percent. In the configure->engine dialog screen you can set the bore, stroke, rod length and clearance. The rod length to stroke deturmins the relation of degrees to cutoff percent. What you calculated is the max angle 5% cutoff is 25.8 degree with a 5" stroke requires a 338" conecting rod. That is the value with a very long conecting rod compared to the stroke. 5" stroke to a 338" conecting rod. at 77" 25.4 and at 5.10 inches 21.2 degrees.
The expansion ratio of an IC engine is the (clearance/clearance+displacement) The valve timming of an IC engine is not involved in expansion as is the steam engine.
It really hard to get better then 5% clearance in an engine. Thats like 21:1 compression ratio in an IC engine. Say you have 5%clearance and 5% cutoff making the expansion 105/10 for a 10.5:1 expansion ratio. Not all that efficient at 5%.
I remember the problems we had getting to an 18:1 compression ration in a motocycle we were running on methanal. Lots of tinkering with piston profiles to clear valves and head. Over reving that engine resulted with a valve compression welded in the top of a piston. Great paper weight and a reminder to never let Rudy on it again. Rudy was 7 at the time and didn't shift. It run at 10000+ for around 10 min before it failed. It could rev out to 16000 RPM with significant power. At 1000 PSIA at takes around 27:1 to expand it down close to 14.696 PSIA atmospheric pressure. That more like 1% cutoff and 3% clearance in a single expansion. The head designs I have been playing with are hard to get under 3% clearance. And what about water slugs on low clearance engines.
My electric valves are limited to around 0.004 seconds minum open duration. So I am very conserned about open duration.
Andy
I'm not sure what you are saying on RPM there.
The way I see it is, as a time function of the valve movement. The fact that an IC valve only operates every other revolution is not significant. The rotating cam at half speed of engine shaft is the only differance. The valves lift and close within so many degreas of rotation. Every revoloution or every other revoloution the timming is tied to the events revoloution. There harmonics I suppose effected by the period of the event. I am not sure where you got 120 degrees. An IC valve open before TDC and closes after TDC a full 180+ degrees.
Your calculation of degrees to cutoff is wrong for a normal crank engine. I don't have time to find the formula right now. It is a bit more complex though. My steam cycle calculator has it built in. Gives cutoff in degrees and percent. In the configure->engine dialog screen you can set the bore, stroke, rod length and clearance. The rod length to stroke deturmins the relation of degrees to cutoff percent. What you calculated is the max angle 5% cutoff is 25.8 degree with a 5" stroke requires a 338" conecting rod. That is the value with a very long conecting rod compared to the stroke. 5" stroke to a 338" conecting rod. at 77" 25.4 and at 5.10 inches 21.2 degrees.
The expansion ratio of an IC engine is the (clearance/clearance+displacement) The valve timming of an IC engine is not involved in expansion as is the steam engine.
It really hard to get better then 5% clearance in an engine. Thats like 21:1 compression ratio in an IC engine. Say you have 5%clearance and 5% cutoff making the expansion 105/10 for a 10.5:1 expansion ratio. Not all that efficient at 5%.
I remember the problems we had getting to an 18:1 compression ration in a motocycle we were running on methanal. Lots of tinkering with piston profiles to clear valves and head. Over reving that engine resulted with a valve compression welded in the top of a piston. Great paper weight and a reminder to never let Rudy on it again. Rudy was 7 at the time and didn't shift. It run at 10000+ for around 10 min before it failed. It could rev out to 16000 RPM with significant power. At 1000 PSIA at takes around 27:1 to expand it down close to 14.696 PSIA atmospheric pressure. That more like 1% cutoff and 3% clearance in a single expansion. The head designs I have been playing with are hard to get under 3% clearance. And what about water slugs on low clearance engines.
My electric valves are limited to around 0.004 seconds minum open duration. So I am very conserned about open duration.
Andy
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Re: Williams ws. Rankin December 06, 2005 09:02PM |
Registered: 18 years ago Posts: 1,249 |
Hello Andy,
On RPM; The valve timeing with a cam is a function of the RPM, higher RPM quicker timeing. The velocity changes imparted to the valve train, lifter, rocker, valve, spring, etc. are also a direct function of the RPM. The forces to move the valve train are due to the acceleration imparted by the cam. Acceleration is directly related to the change in velocity squared. So an engine running at 9,000 RPM would have nine times as much force on its valve train than it would running at 3,000 RPM.
I was using 120 degrees as an example to allow for some dwell at the top of the cam profile, thinking that there must be some, but maybe not as much as 60 degrees to add up to the 180. Just mentioning that there is room to run the valve train twice as many times as the IC engine actually does.
I know the formula would be off a bit for a normal engine and that connecting rod length would have to be figured in. I thought it should be reasonably close though and it's dead on for mine.
Getting less than 5% clearance does take a bit of doing. I am going for around 4% second stage and havn't figured the minimum necessary for the first yet. Some of the smaller diesels can get that. Anyway less clearance is too dependent upon thermal expansions and other variables to be dependable, even with a sponge on top of the piston.
Well back at it. ----------- Bill G.
On RPM; The valve timeing with a cam is a function of the RPM, higher RPM quicker timeing. The velocity changes imparted to the valve train, lifter, rocker, valve, spring, etc. are also a direct function of the RPM. The forces to move the valve train are due to the acceleration imparted by the cam. Acceleration is directly related to the change in velocity squared. So an engine running at 9,000 RPM would have nine times as much force on its valve train than it would running at 3,000 RPM.
I was using 120 degrees as an example to allow for some dwell at the top of the cam profile, thinking that there must be some, but maybe not as much as 60 degrees to add up to the 180. Just mentioning that there is room to run the valve train twice as many times as the IC engine actually does.
I know the formula would be off a bit for a normal engine and that connecting rod length would have to be figured in. I thought it should be reasonably close though and it's dead on for mine.
Getting less than 5% clearance does take a bit of doing. I am going for around 4% second stage and havn't figured the minimum necessary for the first yet. Some of the smaller diesels can get that. Anyway less clearance is too dependent upon thermal expansions and other variables to be dependable, even with a sponge on top of the piston.
Well back at it. ----------- Bill G.
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Re: Williams ws. Rankin December 07, 2005 03:14PM |
Registered: 18 years ago Posts: 1,038 |
Hi Bill
F = M a .. force, mass times acceleration
a = d(v)/dt .. accelerationm, rate of change of velocity with respect to time
v = d(d)/dt .. velocity, rate of change of distance with respect to time
I know all that.
Your 120 degrees was confusing. I thought you were talking about the valve durection from start of opening to it's close. When we were hopping up motocycles we tried two types of performance cams. A Trials cam and a receing cam. The trials cam had steap ramps and a great deal of dwell while the racing cam had long ramps and hard any dwell. With the racing cam we got 16000 RPM out of the engines and the trials cam designed for lowend torque was just RPM crippled. With heavy springs we managed to get 9000 RPM. Valves float problems and clearance problems limited compression ratios to 14:1 with the trials cam. Of course steam inlet valves not protruding into the cylander on opening would not have the clearance problems we had with the steap slope cam.
The IFC-67 steam property formulations do not have enough temperature range to analze supper critical cycles like Harry's using. It is compression that is the problem. The upper limit of 1483F will handle most casses except compression somes times gets the temperature above 1483. But I think suppercritical is a possable soloution to the power range problem I have been talking about. But there is the problem of getting high expansion in a single stage. You say you can get down to maybe 4% clearance. That limits you to 26:1 expansion at 0 cutoff (no output power) As soon as you have any cutoff the expansion ratio is going down. 1% cutoff 20.8:1 expansion, 2% cutoff 17.333:1, 3% cutoff 14.857:1, 4% cutoff 13:1, 5% cutoff 11.556. A multistage engine gets around this problem by dividing up the expansion across 2 or more stages. Using 3 stages of 9:1 expansion for example results in a pressure drop greater then a 27:1 single expansion as there is some pressure drop between stages.
The thing is that we need a range of power. If we don't wont to have a huge transmission we need a huge power range. We need an engine that can produce 0 power up to that required for highway crousing with head room for passing etc.
Say 1 MPH to 85 MPH as a target speed range. Thats a 7225:1 torque range. In my design I splite that up into throttle controled and cutoff control with a mixed transition. Sense the low pow cutoff is using vary short cutoff I am conserned with rough engine running. So I, as a first atempt, use 20 MPH as the lowest speed for pure cutoff control. And pure throttling at 0 RPM transitioning to cutoff at 20 MPH. So cutoff control is over a range of 20 to 85. Which is a 18.06:1 torque range. I still use cutoff controle above 85 holding clearance constant so above 85 expansion is droping off.
With respect to valve open duration someone noted the differance of 15 PSIA air(in an IC) to HP steam. Well we are talking about high compression engines. There is possably little differance in the pressure diferantial across the valves of an IC and high compression steam engine. There should initially be very little pressure differance when the valve opens. In both it's the increassing volume lowing the pressure on the cylander side of the valve that creates the pressure differance. Hard to say exactly what that differance may be though.
Andy
F = M a .. force, mass times acceleration
a = d(v)/dt .. accelerationm, rate of change of velocity with respect to time
v = d(d)/dt .. velocity, rate of change of distance with respect to time
I know all that.
Your 120 degrees was confusing. I thought you were talking about the valve durection from start of opening to it's close. When we were hopping up motocycles we tried two types of performance cams. A Trials cam and a receing cam. The trials cam had steap ramps and a great deal of dwell while the racing cam had long ramps and hard any dwell. With the racing cam we got 16000 RPM out of the engines and the trials cam designed for lowend torque was just RPM crippled. With heavy springs we managed to get 9000 RPM. Valves float problems and clearance problems limited compression ratios to 14:1 with the trials cam. Of course steam inlet valves not protruding into the cylander on opening would not have the clearance problems we had with the steap slope cam.
The IFC-67 steam property formulations do not have enough temperature range to analze supper critical cycles like Harry's using. It is compression that is the problem. The upper limit of 1483F will handle most casses except compression somes times gets the temperature above 1483. But I think suppercritical is a possable soloution to the power range problem I have been talking about. But there is the problem of getting high expansion in a single stage. You say you can get down to maybe 4% clearance. That limits you to 26:1 expansion at 0 cutoff (no output power) As soon as you have any cutoff the expansion ratio is going down. 1% cutoff 20.8:1 expansion, 2% cutoff 17.333:1, 3% cutoff 14.857:1, 4% cutoff 13:1, 5% cutoff 11.556. A multistage engine gets around this problem by dividing up the expansion across 2 or more stages. Using 3 stages of 9:1 expansion for example results in a pressure drop greater then a 27:1 single expansion as there is some pressure drop between stages.
The thing is that we need a range of power. If we don't wont to have a huge transmission we need a huge power range. We need an engine that can produce 0 power up to that required for highway crousing with head room for passing etc.
Say 1 MPH to 85 MPH as a target speed range. Thats a 7225:1 torque range. In my design I splite that up into throttle controled and cutoff control with a mixed transition. Sense the low pow cutoff is using vary short cutoff I am conserned with rough engine running. So I, as a first atempt, use 20 MPH as the lowest speed for pure cutoff control. And pure throttling at 0 RPM transitioning to cutoff at 20 MPH. So cutoff control is over a range of 20 to 85. Which is a 18.06:1 torque range. I still use cutoff controle above 85 holding clearance constant so above 85 expansion is droping off.
With respect to valve open duration someone noted the differance of 15 PSIA air(in an IC) to HP steam. Well we are talking about high compression engines. There is possably little differance in the pressure diferantial across the valves of an IC and high compression steam engine. There should initially be very little pressure differance when the valve opens. In both it's the increassing volume lowing the pressure on the cylander side of the valve that creates the pressure differance. Hard to say exactly what that differance may be though.
Andy
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Re: Williams ws. Rankin December 07, 2005 07:05PM |
Registered: 18 years ago Posts: 1,249 |
Andy,
I found a free addin for excel which also works in Open offices calc sheet if one doesn't have excel.
I got off onto the cam design and havn't had time to play with it much other than to see if it worked. It uses the IFC-97 tables and goes up to around 1800 deg. if one converts stuff from metric. Saves a couple hundred bucks.
[www.cheresources.com]
Hope it can be of use.
It sounds as if you have more development time on your engine than I have on mine. I havn't gotten into torque ranges yet, but discarding port and valve throttling (dangerous to do) mine has a torque range that is set much by the cutoff. Like yours though I was planning some throttling at long cutoff in the lower RPMs. Also possibly some goosing of the second stage to get off the line.
With six first stage and six second stage pistons rough running shouldn't be too big a problem. A smaller version of the engine could get by with four & four. The minimum expansion at 30% cutoff in my first stage is kept rather low, between 3-4 but this keeps the maximum pressure going into the second stage just below 300 psia. and the temps below 900. That way the second stage can have a hopefully unused popoff valve because the pistons are too big to slam with a thousand pounds pressure. Trying to keep them as light as possible and to take a thousand pounds they would have to weigh nine pounds apiece.
Got a question for you. If (first stage) we recompress back into the steamchest at inlet pressure and our temperature difference isn't too great, say by use of a check valve, is there much more or any loss of efficiency compaired to recompressing into an extra clearance space? It seems offhand that disregarding temperature mixing all the work of recompression would be returned in both methods.
Thank You ---------- Bill G.
I found a free addin for excel which also works in Open offices calc sheet if one doesn't have excel.
I got off onto the cam design and havn't had time to play with it much other than to see if it worked. It uses the IFC-97 tables and goes up to around 1800 deg. if one converts stuff from metric. Saves a couple hundred bucks.
[www.cheresources.com]
Hope it can be of use.
It sounds as if you have more development time on your engine than I have on mine. I havn't gotten into torque ranges yet, but discarding port and valve throttling (dangerous to do) mine has a torque range that is set much by the cutoff. Like yours though I was planning some throttling at long cutoff in the lower RPMs. Also possibly some goosing of the second stage to get off the line.
With six first stage and six second stage pistons rough running shouldn't be too big a problem. A smaller version of the engine could get by with four & four. The minimum expansion at 30% cutoff in my first stage is kept rather low, between 3-4 but this keeps the maximum pressure going into the second stage just below 300 psia. and the temps below 900. That way the second stage can have a hopefully unused popoff valve because the pistons are too big to slam with a thousand pounds pressure. Trying to keep them as light as possible and to take a thousand pounds they would have to weigh nine pounds apiece.
Got a question for you. If (first stage) we recompress back into the steamchest at inlet pressure and our temperature difference isn't too great, say by use of a check valve, is there much more or any loss of efficiency compaired to recompressing into an extra clearance space? It seems offhand that disregarding temperature mixing all the work of recompression would be returned in both methods.
Thank You ---------- Bill G.
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Re: Williams ws. Rankin December 08, 2005 10:16AM |
Registered: 18 years ago Posts: 318 |
Jim,
There is some debate in the Williams family as to when to release the notes. So, I don't know when they become available.
Bill, You should see the intake springs on the Williams engine. You cannot compress them with an off-the-shelf compressor. You need to make a special one. The Williams is 16 expansions for optimum. I believe the last few expansions are more trouble than they are worth. Expanding out to the last pound is a mistake in my opinion.
Howard
--
There is some debate in the Williams family as to when to release the notes. So, I don't know when they become available.
Bill, You should see the intake springs on the Williams engine. You cannot compress them with an off-the-shelf compressor. You need to make a special one. The Williams is 16 expansions for optimum. I believe the last few expansions are more trouble than they are worth. Expanding out to the last pound is a mistake in my opinion.
Howard
--
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Re: Williams ws. Rankin December 08, 2005 12:04PM |
Registered: 18 years ago Posts: 1,249 |
Hello Howard, Good to hear from You again.
I was thinking that heavy springs would be involved. I wonder about the long term wear from using such heavy springs.
I hope the Williams family, if they are contemplating releasing the notes, do so while there are still people around that care and could benefit from them. Abner Dobles papers for instance are quite interesting and a valuable resource for ideas. Or else throw some money at further development.
Agreed that expanding out to the last pound could be more loss than gains. I do think, however, that we can profitably get closer than we have been able to do with a single stage engine. The compound is the only piston way to do that, and, I believe, the best way to profitably start heading into higher temperatures and pressures where the greater expansions are necessary.
I am not in any way disregarding Jims idea of using a turbine for the second stage expander. I see that as the "other" valid alternative and believe it really needs to be explored.
At any rate the compound is flexable and cylinder sizes can be adjusted while in the test stand phase if the big expandions are found to be unprofitable.
This "modern engine" is doable, and I believe the Williams have helped to point the way, but we must take it from there. I think the compound Williams is the next step, hope so.
Best of days to Everyone ------------- Bill G.
I was thinking that heavy springs would be involved. I wonder about the long term wear from using such heavy springs.
I hope the Williams family, if they are contemplating releasing the notes, do so while there are still people around that care and could benefit from them. Abner Dobles papers for instance are quite interesting and a valuable resource for ideas. Or else throw some money at further development.
Agreed that expanding out to the last pound could be more loss than gains. I do think, however, that we can profitably get closer than we have been able to do with a single stage engine. The compound is the only piston way to do that, and, I believe, the best way to profitably start heading into higher temperatures and pressures where the greater expansions are necessary.
I am not in any way disregarding Jims idea of using a turbine for the second stage expander. I see that as the "other" valid alternative and believe it really needs to be explored.
At any rate the compound is flexable and cylinder sizes can be adjusted while in the test stand phase if the big expandions are found to be unprofitable.
This "modern engine" is doable, and I believe the Williams have helped to point the way, but we must take it from there. I think the compound Williams is the next step, hope so.
Best of days to Everyone ------------- Bill G.
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Re: Williams ws. Rankin December 08, 2005 12:34PM |
Registered: 18 years ago Posts: 1,441 |
Howard & Bill,
At least now I know who is sitting on the Williams data, the family.
Now you see the big problem with heavy valve springs. essential with short cutoff. You simply must use wide roller followers and wide cam lobes to get any kind of life out of them. Such a valve gear begins to absorb more power than I like to see.
This is why, no matter how I approach this new engine problem, I keep coming back to not giving a really short cutoff in the engine itself; but compound with a turbine.
Again, because of balance, heat and flow losses, and size considerations.
I still cannot abandon the three rotor Wankel, because of packaging ease and really big "Piston" area, coupled with a radial inflow turbine. BUT; with a variable nozzle ring.
JC
At least now I know who is sitting on the Williams data, the family.
Now you see the big problem with heavy valve springs. essential with short cutoff. You simply must use wide roller followers and wide cam lobes to get any kind of life out of them. Such a valve gear begins to absorb more power than I like to see.
This is why, no matter how I approach this new engine problem, I keep coming back to not giving a really short cutoff in the engine itself; but compound with a turbine.
Again, because of balance, heat and flow losses, and size considerations.
I still cannot abandon the three rotor Wankel, because of packaging ease and really big "Piston" area, coupled with a radial inflow turbine. BUT; with a variable nozzle ring.
JC
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Re: Williams ws. Rankin December 08, 2005 12:49PM |
Registered: 18 years ago Posts: 835 |
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Re: Williams ws. Rankin December 08, 2005 06:54PM |
Registered: 18 years ago Posts: 1,441 |
Harry,
You could; but often in laying out a system for a car, the plumbing becomes a nightmare. Only seems doable if the engine and steam generator are right next to each other. Otherwise the losses become just too great.
Also, controlling the reheat temperature, which Doble never bothered to do, requires at least a second normalizer. I think the extra weight and complication would far outweigh the little gain.
JC
You could; but often in laying out a system for a car, the plumbing becomes a nightmare. Only seems doable if the engine and steam generator are right next to each other. Otherwise the losses become just too great.
Also, controlling the reheat temperature, which Doble never bothered to do, requires at least a second normalizer. I think the extra weight and complication would far outweigh the little gain.
JC
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Re: Williams ws. Rankin December 09, 2005 01:16PM |
Registered: 18 years ago Posts: 1,249 |
Harry,
I keep thinking of that marvelous six cylinder radial of yours exhausting close and directly into the turbine nozzle ring, or six individual adjustable nozzles. Little to no recievers and keep the pressure drop from exhaust to nozzle as minimal as possible. Even with bump valves starting at the pressures and temps your engine runs at, throttling should be efficient over a five to one power range.
It should be a high RPM screamer with efficiencys way up there.
That radial is just perfectly shaped to connect to a turbine. That really was the "other" option I was thinking about.
Jims mentioning of the Wankle keeps comming up as something also hard to let go of. I don't see the Wankle as a good first or high pressure stage but as an exellent second stage possibility where the temperature and clearance problems won't get in the way. Am wondering about the Wankles abilities to do recompression at lower second (or third) stage pressures.
The cam designing is comming along just great. I am awaiting the arrival of a catalog from Timkin, they must have sent it by camel. They make a line of cam and rocker rollers for engines, and I havn't run across other manufacturers that mention it. Do any of you know of other contacts for these items?
Best to Everyone ------------ Bill G.
Howard, with the low cam lift and short cutoff, how much valve lash does that Williams engine use? It seems that with lifts in the area of .100 that the standard lash of around .050 would unmanagable.
Edited 1 time(s). Last edit at 12/11/2005 10:13AM by Bill Gatlin.
I keep thinking of that marvelous six cylinder radial of yours exhausting close and directly into the turbine nozzle ring, or six individual adjustable nozzles. Little to no recievers and keep the pressure drop from exhaust to nozzle as minimal as possible. Even with bump valves starting at the pressures and temps your engine runs at, throttling should be efficient over a five to one power range.
It should be a high RPM screamer with efficiencys way up there.
That radial is just perfectly shaped to connect to a turbine. That really was the "other" option I was thinking about.
Jims mentioning of the Wankle keeps comming up as something also hard to let go of. I don't see the Wankle as a good first or high pressure stage but as an exellent second stage possibility where the temperature and clearance problems won't get in the way. Am wondering about the Wankles abilities to do recompression at lower second (or third) stage pressures.
The cam designing is comming along just great. I am awaiting the arrival of a catalog from Timkin, they must have sent it by camel. They make a line of cam and rocker rollers for engines, and I havn't run across other manufacturers that mention it. Do any of you know of other contacts for these items?
Best to Everyone ------------ Bill G.
Howard, with the low cam lift and short cutoff, how much valve lash does that Williams engine use? It seems that with lifts in the area of .100 that the standard lash of around .050 would unmanagable.
Edited 1 time(s). Last edit at 12/11/2005 10:13AM by Bill Gatlin.
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Re: Williams ws. Rankin December 12, 2005 11:06AM |
Registered: 18 years ago Posts: 1,038 |
Hi Bill
One thing, with over compressing into the steam chest, is that you are doing constant pressure work forcing the steam through the valve into the steam chest. In an ideal cycle we are not including flow losses. But in a real engine there must be some pressure differance to force the flow. So there is some throttling loss both ways that would increase over all losses. How much? Better or worse then varing clearance so as to not over compress? Hard to answer these questions.
The IFC-97 formulations should go up to 3632F below 1450 PSIA. Above 1450 PSIA the upper temperaturelimit is 1472F.
273.15 K <= T <= 1073.15 K p £ 100 MPa
1073.15 K <= T <= 2273.15 K p £ 10 MPa
See region 5 of the IFC-97 formulations [www.iapws.org]. Maine IAPWS page: [www.iapws.org]
Andy
One thing, with over compressing into the steam chest, is that you are doing constant pressure work forcing the steam through the valve into the steam chest. In an ideal cycle we are not including flow losses. But in a real engine there must be some pressure differance to force the flow. So there is some throttling loss both ways that would increase over all losses. How much? Better or worse then varing clearance so as to not over compress? Hard to answer these questions.
The IFC-97 formulations should go up to 3632F below 1450 PSIA. Above 1450 PSIA the upper temperaturelimit is 1472F.
273.15 K <= T <= 1073.15 K p £ 100 MPa
1073.15 K <= T <= 2273.15 K p £ 10 MPa
See region 5 of the IFC-97 formulations [www.iapws.org]. Maine IAPWS page: [www.iapws.org]
Andy
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Re: Williams ws. Rankin December 12, 2005 11:17AM |
Registered: 18 years ago Posts: 1,038 |
Hi Howard.
The williams got 16:1 expansion? At 4% cutoff you would have to have 2.4% clearance to gat that 16:1. Sounds questionable. Are you sure about that? And then that is figureing a full stroke expansion. The uniflow ports reduce the expansion ratio.
As for trying to get close to full expansion. There is the problem, in a throttled engine, of reducing the power range to nil or lossing efficiency over expanding, below exhaust pressure, when throttling the power down.
Andy
The williams got 16:1 expansion? At 4% cutoff you would have to have 2.4% clearance to gat that 16:1. Sounds questionable. Are you sure about that? And then that is figureing a full stroke expansion. The uniflow ports reduce the expansion ratio.
As for trying to get close to full expansion. There is the problem, in a throttled engine, of reducing the power range to nil or lossing efficiency over expanding, below exhaust pressure, when throttling the power down.
Andy
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Re: Williams ws. Rankin December 12, 2005 01:02PM |
Registered: 18 years ago Posts: 835 |
Hi Bill,
The 6cyl engine has a form of a popet valve only it is not, it is preasure balanced for the high preasure as springs would be hurendous. The 2cyl is a bump valve and is only for one speed it is a dumd engine and some times although it is simple more control would be nice. I like a radial engine as it balances very easialy. Has one through and two main bearings and one cam.
Jim, Have you looked at a radial configuration as compared to the Wankel, I would like your opinion on this. Maybe as you said the exaust could turn a direct mounted turbine as a second stage? And as Bill suggests each port would line up with the wheel as a nozzel.
Harry
The 6cyl engine has a form of a popet valve only it is not, it is preasure balanced for the high preasure as springs would be hurendous. The 2cyl is a bump valve and is only for one speed it is a dumd engine and some times although it is simple more control would be nice. I like a radial engine as it balances very easialy. Has one through and two main bearings and one cam.
Jim, Have you looked at a radial configuration as compared to the Wankel, I would like your opinion on this. Maybe as you said the exaust could turn a direct mounted turbine as a second stage? And as Bill suggests each port would line up with the wheel as a nozzel.
Harry
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Re: Williams ws. Rankin December 12, 2005 09:27PM |
Registered: 18 years ago Posts: 318 |
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Re: Williams ws. Rankin December 12, 2005 10:39PM |
Registered: 18 years ago Posts: 1,249 |
Harry,
What timeing are you getting with the bump valve? How much lift with it?
I have been kicking around the idea of using a pneumatic spring (steam) for the poppet valve. This way the spring pressure could be automatically adjusted to the steamchest pressure, and as RPMs go up the spring rate would stiffen while still maintaining the same seat pressure. At 1000 psia inlet I would need 110pounds of seat force to equal recompression pressure. That's about what a standard valve spring is at.
Such pneumatic springs have been mentioned before for racing engines but are they a hassle as far as reliability. How are they sealed?
Andy, thanks for going over the recompression thing again. Even though we came up with a nifty variable clearance mechanism, I'm running into space to put it considerations and am examining a possible alternative of recompressing back into the steamchest/port.
I guess the only way to get to the high pressure/temperature area that isn't available with the IFC-97 would be to calculate from the available data as near as it goes and extrapolate an appropriate polytropic expression. (k) The calculated results would have to be accurate within any mechanical variances of a device.
Thank You -------- Bill G.
What timeing are you getting with the bump valve? How much lift with it?
I have been kicking around the idea of using a pneumatic spring (steam) for the poppet valve. This way the spring pressure could be automatically adjusted to the steamchest pressure, and as RPMs go up the spring rate would stiffen while still maintaining the same seat pressure. At 1000 psia inlet I would need 110pounds of seat force to equal recompression pressure. That's about what a standard valve spring is at.
Such pneumatic springs have been mentioned before for racing engines but are they a hassle as far as reliability. How are they sealed?
Andy, thanks for going over the recompression thing again. Even though we came up with a nifty variable clearance mechanism, I'm running into space to put it considerations and am examining a possible alternative of recompressing back into the steamchest/port.
I guess the only way to get to the high pressure/temperature area that isn't available with the IFC-97 would be to calculate from the available data as near as it goes and extrapolate an appropriate polytropic expression. (k) The calculated results would have to be accurate within any mechanical variances of a device.
Thank You -------- Bill G.
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Re: Williams ws. Rankin December 13, 2005 09:50AM |
Registered: 18 years ago Posts: 1,441 |
Bill,
Pneumatic valve springs are used primarily in Formula One, because those engines now go to 20K rpm. They are completely reliable. They use nitrogen at about 150 psi in a sealed system. Sealed with both rings or a rolling diaphram, depending on which team you are talking about. The rules forbid them in Indy car, IRL, and NASCAR racing.
However, those F-1 engines have a life span of some ten hours before a complete rebuild.
There are lots of papers on these pneumatic springs in the SAE publications and in several of the newer books on race car engineering.
JC
Pneumatic valve springs are used primarily in Formula One, because those engines now go to 20K rpm. They are completely reliable. They use nitrogen at about 150 psi in a sealed system. Sealed with both rings or a rolling diaphram, depending on which team you are talking about. The rules forbid them in Indy car, IRL, and NASCAR racing.
However, those F-1 engines have a life span of some ten hours before a complete rebuild.
There are lots of papers on these pneumatic springs in the SAE publications and in several of the newer books on race car engineering.
JC
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