Williams ws. Rankin

Hi Bill,
The timing on a bump valve is same open as same close. We are useing 5% up to 6.65% however it has a starting problem as you decrese clearance volume. it likes high preasure as it will not run below compression preassure. In the cyclone it has to have a reconpression tube and with this it will start easily and run on as little as 75psi. Yesterday we ran some more test with the tube removed and the engine ran not nearly as well. I belive Jay Carter and Bill Ryan are running at about an 8 to 1 compression ratio where this engine is happy in normal conditions.
Harry

Hi Bill

The IAPWS-95 scientific formulations cover higher pressure and temperature ranges then the IFC formulations and are more accurate. The cost is computation time and there are no backword formual to speed it up. You are stuck with temperature and density as input. Other properties are calculated. So if you have entropy and specific volume you have to solve for density and temperature. And the saturation line is a bitch. The 95 formulations will sometime cause floating point errors if you give them a temperature and density of a mixture. Have to check if it is with in the mixture region and solve for the saturation pressure. Still working on it.

Andy




Edited 1 time(s). Last edit at 12/13/2005 04:51PM by Andy.

Hi Howard

"Andy remember the William is a O clearance engine " Havn't heard that clame before. Hard to believe.

Are you confusing real zero clearance with efective zero clearance. The term "efective zero clearance" that Jerry and I use means that the compression cycle eliminating clearance loss makes efficiency the same as the zero clearance text book cycle.

Not the same as real zero clearance when figuring expansion ratio. The residual steam in the clearanc and inlet steam at cutoff are bothe expanded. Efective zero clearance does not effect the expansion ratio calculation. Efective zero clearance is saying that the losses due to that clearance are mostly eliminated. But the expansion ratio is the change in volumes from cutoff to end of expansion and must include the real clearance volume.

If it had real zero clearance you could not use compression. Zero clearance would make for infinite compression ratio.

I use:

compression ratio = (clerance + x)/clearance.

x is percentage of displacement remaining when exhaust closes.

Expansion ratio = (clearane+1)/(clearance + cutoff)


To explaine "effective zero clearance", using the expansion ratio formula above, we make the zero clearance cycle expansion ratio the same as the non zero clearanc engine with compression back to inlet pressure and the have the same ideal cycle efficincy for full expansion cycles. The compression makes the ideal cycle efficiencies equal where with out compression the clearance space loss reduces efficincy compared to the ideal zero clearance cycle.

Andy




Edited 1 time(s). Last edit at 12/14/2005 11:39AM by Andy.

HI Bill your lucky your stuff comes by camel mine comes by pony express

Thanks Everyone.

Yea, still waiting for the catalog. I always hope big companys take me seriously enough to send stuff. Once I'm on a problem my head doesn't quit untill I've found answers and chosen an ultimate. Sometimes it just seems to go on without me. LOL

I am getting the cam geometry down and soon will need real world roller dimentions and load ratings to continue. Does anyone know what kind of timeing they were getting with the Skinner duel cam system?

Good days to All ------------ Bill G.

Hi Andy,

I assumed you knew I meant zero clearance by compression. As we all know there is no such thing as zero physical clearance. Also, you have to remember the steam filling the clearance volume is recycled steam. So the inlet steam does not have to fill any clearance volume.

You seem to have a thing against the Williams engines. How many steam cars do you know that can cruse at 85 mph with an 8 gallon fire, with the fire shutting off, and especially ones that are less streamlined than an old jeep and can top out at 100 mph.

Actually you and Jim don't know this, but there was a recent test done on the Williams engine by a highly prestigious testing institute. We are suppressing the test results as we find the results suspicious -- they were too good in our view. We don't think the testing institute know what they were doing but they certainly knew how to charge for their test. That is why Tom is redoing the test. I am operating under a non-disclosure agreement with the Williams family, so there are certain things I cannot talk about or talk around with out giving too much detail.

I don't see how it is fair for you to keep comparing a paper engine with an actual operating engine. Actually I can't see how you can compare your engine even with a Newcomen. The Newcomen was proven technology and a commercial success for its day even if you had to own an coal mine to operate one.

I don't mean to be argumentative but this is the way I see it.

Howard




.

Howard, Andy,
There is only one way this entire question about the excellence, or not, of the Williams engine is going to be settled.
Secret tests by some lab that does not know what to look for and how to test the engine, is only going to add to the confusion. And yes, I knew about the supposed test; but not the details.
Quite frankly, I was not interested enough to find out, because minimal clearance and high compression are well established design criteria to enhance the efficiency of any steam engine. A test would only verify this AGAIN. This concept was established and proven over 80 years ago for heaven's sake.

Indeed, Williams built a good engine; but for once I would love to know what secret ingredient the Williams engine had that was any different than a few really good engines had in the past. To date, I only see that they applied what was needed to get a low water rate, in their viewpoint and at that time. Concepts that most steam car engine builders did not use for unknown reasons. Probably ignorance.
There are a number of items one simply has to have when designing a steam engine to minimize losses. High compression and minimum clearance volume are only two of them, and the rest are just as important.

When the Williams engine is tested on a dynomometer, with full description of ALL the test procedures, proof of calibration of ALL the test equipment, with the dyno's constant included, which a lot of people don't know about and ignore, then we all will be able to evaluate the Williams engine correctly.
Until then it is still pure conjecture and guessing.

Let us all wait until Tom sets up a real test system, run the Williams engine at correct pressure and temperature, and have enough verified instrumentation so that any guessing, or worse, extrapolation of data is eliminated, and then publish the test results.

A steam car that shows an average of 8-9 gph for an hour with the fire going on and off, can cruise at 80 mph and go well over 100 mph? I know of at least three good Series E Dobles with draft boosters that do this.
JC

On the Williams engine are they trying to prove a low water rate or fuel consumption in lbs per hp per hr. If so the the boiler efficency would come into play as the total package is what counts. I belive, correct me I am wrong but high comp uniflow 27/1 comp bottoms out on water rate in the 6 lb hp range. Then add acc. and boiler for the total.
Harry

Hi Howard

Sorry the word "confusing" wasn't ment to be insulting. But efective zero clearance isn't real zero clearance And when figuring real expansion ratio we need to figure the volume change. You said "The Williams is 16 expansions for optimum." From the Williams data I have the shortest cutoff is 4% so the clearance would have to be like 2.4% probably much less to efect a 16:1 expansion. It is a uniflow engine so the exhaust opens quite a bit before TDC. Don;t have the exhaust points in my Williams info. You don't have full stroke of expansion available. But there would be some partial expansion during that (exhaust) part of the stroke. So I stll say 16:1 sounds a bit high. It has nothing to do with what I think about the Williams engine. That was my argument to your clame of 16:1 expansion ratio. I just find it very hard that engine could have a 2.4% minum clearance to get that expansion ratio. I just figured you had a momentary slip or something. I do the same. Maybe you could simply explaine how you figured the Williams to have a 16:1 expansion ratio.

I do take exception to calling it a Williams cycle. From all I have heard they couldn't figure cycles. It's not an honor they deserve.

I am not a fan of uniflow engins for variable power. What I have found through analysis does show posable other reasions that the uniflow engine is so efficient. Compression can improve efficiency by a great deal. An ideal compression cycle compared to a non-compression cycle shows efficiency differances equal or greater then what a simular uniflow does. I admit that I have data for only a few examples. Hard to find.

It isn't that I don't like the uniflow Williams. It isn't all that complex. But I beleve that high compression can be applied to a counterflow with better compression.

The Williams is limited by not having exhaust timming controll. I only have paper idea. But that does not nagate the result. My idea quite different then the Williams engine but utilizes the same principals. I do not doubt the results of Williams tests. But on the other hand I think it workd in a nerrow power band. It would have to be limited as it uses fixed cutoff points and throttling controll. My idea is to use cutoff control with no throttling over as much as 90% of the normal power range. The expansion ratio is kept constant over that part of the power range. What I am talking can not be done with a uniflow. The exhaust close point must be varied to keep the compression to just the inlet pressure. The clearance and cutoff are varied to change power while keeping a constant end of expansion presure. What I am doing is varing the relation between residual(recycled in the engine) steam and fresh inlet steam.

I have did a lot of analysis of the Williams engine. I origionally wrote my SCAR steam cycle analysis program to investigate high cmpression cycles. That was around 1994. Have been an avocate of high compression every sense.

The clearance loss is inversly dependent on cutoff. The short the cutoff (constant clearance) the more the clearance loss. The steam it takes to fill the clearance represents a power loss equal to it's voume times pressure. It not hard to get a 15 to 10 percent loss due to clearance at high expansion ratios. Filling that space with compressed residual steam eliminates that loss at a cost of power reduction. Ideally the compresson work done to the residual steam is returned upon expansion. There is some debate as to the state of the steam at the start of compression. In the worst case it would be as if it fully expanded to exhaust pressure isentropicly. In the best it would expand isoenthalpy thus having higher heat content. Take the isentropic case(lowest binifit). It entropy would be the same as at the start of expansion and would compress to inlet pressure and temperature. Upon compression it would be identical to the inlet steam. In the isoenthalip case it would have initial higher heat content and pon compressio to inlet pressure would be at a higher temperature then the inlet steam(bet case). I beleve it to hae higher heat content then the isentropic expansion. Can't say it is isoenthalip expansion though either. But the chart posted on another by Peter hide clearly shows compression temperature raises above the inlet temperature.

[www.steamautomobile.com]

Most believe the idea of cooling condensation and reheating of the inlet port to account for uniflow high efficiency. But my compression cycle analsis show equilivent efficiency gaines due to compression. That chart shows the temperature increasse above inlet temp. I can get very close to the temperatures on that chart with my scar program. compression to 700+ degrees no mater what resionable pressure I use and a temperature of 470F. My scar program uses the isoenthalip method. To be clear here. I am talking about the expansion that takes place when the exhaust opens. From end of expansion to exhaust pressure. Ted Pritchard brought this up on another thread. said we wrong to use constant enthalpy. My books say that engine valves are throttling processes. That is what I used in my scar program. However when you apply the conservation of energy law a constant enthalpy can't be wight. I can't get an energy balance using isentropic drop either. So either the conservation law shudn't be applied to the constant pressure work. Constant pressure work is like a motor. No differnt then a hydrolic motor. Sense a partial expansion cycle is like a full expansion cycle on top of a constant pressure non-expansion cycle the conseration law only applies to the expansion cycle part?

In any case there is an efficiency gain using compresion. It would be a little less then the clearance loss to more then the clearance loss. My scar program shows compression using the isoenthalip process to bring the efficiency to above the zero clearance non-compression partial expansion cycle but never above what a full expansion cycle would get. In other words compression gains more then would be lost to clearance not compressing.

Andy

Harry,
Quite correct. The efficiency of every single component in the whole cycle enters into the water and fuel rate calculations. Things like: Boiler, feed pump, condenser fan, vacuum pump, electrical load, the auxiliary drive method, efficiencies all enter into this.
Most certainly the engine's water rate is paramount and the lower the better. There are at least a dozen things that influence this, and all have to be optimized before one can get below an 8 lb water rate.THAT is one rough road.

I don't think this Williams stuff should be called a new cycle either. So far, unless some magic appears. What they did was actually nothing new. They put several factors into their engine design at one time; but not inventing anything that had not been thought of individually decades earlier.
JC

Andy,

Recompressing isentropically expanded and exhausted residual steam back to inlet pressure and temperature effects a zero clearance and the recompresssed steam gives up all of the work energy put into it upon expansion like a spring. So the work of recompression all comes back. (ideal here)

The Williams engine claims a further benefit from recompressing the non idealy expanded and exhausted higher than starting entropy steam back to inlet pressure and a much higher temperature again effecting a zero clearance. The recompressed steam then in mixing gives up heat to the inlet steam and looses heat it'self. The total heat then is what the fresh inlet steam had plus what the recompressed steam added.

There is though an irreversable mixing of two temperatures and it doesn't seem that more work would be done by the expansion of the two masses of steam mixed together than there would be by expanding them separately.

The Williams patent said that the benefit was by not dipping into the saturated steam area toward the end of expansion.

Now lets take a rough example from the chart. Operating at an inlet of 200 psia and an entropy of 1.75, about 1420 BTU, around 790 degrees F.

Dropping to an enthalpy of 1160 just above the saturation line. This gives a drop of 260 BTUs. With 33% residual steam the work of compression back to inlet conditions then is about 86 BTUs. Work out then around 174 BTUs.

Ok start the recompression from an entropy of 1.875 and again 1160 BTUs. Our constant enthalpy. Recompressing to 200psia 1580 BTUs and about 1100 deg F. Again 33% residual steam.

Mixing with the inlet steam we have 1580*.33= 521 BTUs plus 1420*.66= 937 for a total of 1458 BTUs. About 870 degrees. Dropping to 1160 is 298 BTUs and the work of recompression is 1580 -1160 = 420 420* .33= 138. So the work out is 298 - 138= 160BTUs

That is a difference of 8% in efficiency with the same percentage of residual steam, different clearance. The mixing of different temperatures is generally a waste of potential work somewhere and needs to be watched closely. In this case it was the potential of an engine that could have operated on the temperature difference of 1100 degrees and 790 degrees.

I still say that the exhaust process is pretty much a constant entropy affair inside the cylinder and an increase in enthalpy from the work done upon it by the residual steam outside the cylinder. Again if we were exhausting through nozzles instead of exhaust ports it would all look pretty much the same from inside the cylinder.

Best to All ------------ Bill G.

HI Bill .you got It right. Howard

Hi Bill

Havn't seen anything like that in the Williams clames. But in Jerry's analysis of the Williams engine, that is the explanation. From what I have heard the Williams were not able to do an analysis. They didn't have the backgrounds.

Jerry Peoples has done a lot of study on the Williams engine. The principals are all in Jerry's paper. The only thing is, that he uses ideal gas formula. Steam is not an ideal gas. Expecially wet steam. The numbers come ot quite a bit different using IFC steam formuaions. That does not in any way invalidate Jerry's paper though. The principals do not change for a real gas.

[www.steamautomobile.com] STUMPF CHART

Look close at the temperature line, in the STUMPF chart atachment, during compression. The inlet temperature is 470 degrees. The compression atains a temperature if 934 as the inlet opens. Seams the exhaust process is not isentropic. There is considerable more heat contained in the steam at the start of compression then if it expanded isentropicly. There are things wrong with that chart though. With a compression temperature of 934 I would expect to see a temperature greater then 470 at cutoff. The mixture temperature at cutoff should be greater then the lower component of the mix. The temperature change during exhaust open is inconsistant between the two engines. That need an explanation. If both were exhaust into like exhaust systems I would expect to see simuler temperature plots on the down stroke. Instead we see a temperature from exhaust open to BDC of a 169 degrees in the counter flow engine and only 47.5 degres in the uniflow. Seems they must have drasticly different exhaust pressure of something. So with that inconsisancy it's hard to say. The counter flow seams to support isentropic expansion while the uniflow seams to suppoer the isenthalip expansion. Which make me sceptical of the chart.

The problem with you cycle example is the phisical implementation in an engine. Couldn't get that cycle in a uniflow with out excessive early exhaust opening. Anyway hear is a comparitive cycle as close as my scar program can get. But it is desined to analyze a high compression counterflow. So no early exhaust open and the exhaust close is settable to any desired possable compression.

Non-compression cycle, Engine clearance 5.33%, cutoff 10.8% Admitted part 95.6% residual part 4.4%

inlet properties: pressure 200 PSIA, temperature 790F, specific volume 3.66038 ft^3/lb, enthalpy 1420.4 BTU/lb, entropy 1.76223
cutoff properties: pressure 200 PSIA, temperature 767F, specific volume 3.59104 ft^3/lb, enthalpy 1409.0 BTU/lb, entropy 1.75302
expansion properties: pressure 17.3 PSIA, temperature 233.9F, specific volume 23.4498 ft^3/lb, enthalpy 1160.1 BTU/lb, entropy 1.75302
exhaust properties: pressure 15.0 PSIA, temperature 232.2F, specific volume 27.0964 ft^3/lb, enthalpy 1160.1.4 BTU/lb, entropy 1.76875
compression properties: pressure 15.0 PSIA, temperature 232.2F, specific volume 27.0964 ft^3/lb, enthalpy 1160.1.4 BTU/lb, entropy 1.76875
Exhaust close at TDC.

compression cycle, Engine clearance 5.33%, cutoff 10.8% Admitted part 68.0% residual part 32.0%

inlet properties: pressure 200 PSIA, temperature 790.0F, specific volume 3.66038 ft^3/lb, enthalpy 1420.4 BTU/lb, entropy 1.76223
cutoff properties: pressure 200 PSIA, temperature 809.0F, specific volume 3.71946 ft^3/lb, enthalpy 1430.1 BTU/lb, entropy 1.76997
expansion properties: pressure 17.4 PSIA, temperature 259.1F, specific volume 24.2883 ft^3/lb, enthalpy 1172.3 BTU/lb, entropy 1.76997
exhaust properties: pressure 15.0 PSIA, temperature 257.6F, specific volume 28.1536 ft^3/lb, enthalpy 1172..3 BTU/lb, entropy 1.78607
compression properties: pressure 200.0 PSIA, temperature 849.3F, specific volume 3.8447 ft^3/lb, enthalpy 1450.1.9 BTU/lb, entropy 1.78607
Exhaust close at 66.7% of return stroke. Compression for 33.7% of stroke. A LONG EXHAUST FOR A UNIFLOW.

What happens at low power. 26.4% Admitted to 73.6% residual. 4.2% cutoff 15% clearance.
inlet properties: pressure 200 PSIA, temperature 790.0F, specific volume 3.66038 ft^3/lb, enthalpy 1420.4 BTU/lb, entropy 1.76223
cutoff properties: pressure 200 PSIA, temperature 1046.6F, specific volume 4.45008 ft^3/lb, enthalpy 1553.6 BTU/lb, entropy 1.85914
expansion properties: pressure 19.9 PSIA, temperature 434.5F, specific volume 26.6542 ft^3/lb, enthalpy 1255.6 BTU/lb, entropy 1.85914
exhaust properties: pressure 15.0 PSIA, temperature 433.3F, specific volume 35.3042 ft^3/lb, enthalpy 1255.6 BTU/lb, entropy 1.88996
compression properties: pressure 200.0 PSIA, temperature 1136.6F, specific volume 4.72402 ft^3/lb, enthalpy 1601.4 BTU/lb, entropy 1.88996
Exhaust close at 2.9% of return stroke. Compression for 97.1% of stroke. A very short exhaust for a uniflow.

Both compression cycles were figured using cutoff and clearance parameters of a non-compressions cycle at 200 PSIA pressure operating from 1420.4 BTU/lb inlet enthalpy to 1160 BTU/lb end of expansion enthalpy after expansion.

Hi Andy,

I'm going to transfer that all to a T_S chart to make more sense of it.

Question: Are there any patents on the Skinner duel cam setup? I need to find out if I've reinvented the wheel here or not. Sometimes one persons "EUREKA!" is another persons "WELL, DUH".

Thank You --------- Bill G.

Bill, Andy,
What with all this calculation, the most important, what is your predicted water rate? From the indicator diagram to the subtraction of the machine( pump, engine friction, ect). You may find the latter to take up gaines in other areas.
what is base water rate that is calculated.
Harry

Bill.i Will ask Tom he nose Hal Fuller. Howard

Hello Harry,

What I was showing was that recompressing to a higher temperature and the same pressure is a loss, and that that loss happens due to the mixing of steam at two different temperatures. The savings of recompressing to a higher temperature is only when it would keep the engine out of the saturated steam area. That because saturated steam is so hard to expand effectively and the engine has little extra expansion capability.

The compounds have expansion capability to spare and can more effectively dip into the wet steam area.

A thought did strike, however. What if much of the clearance space were made of tubing and the tubing were heated to the higher temperature. The temperature of the mix would be higher at no extra cost of recompression work. The valves could then operate at a lower temperature than an equivalent raise in superheat would bring, and if neccessary the cylinder walls could be cooled with recirculated water like the ultimax.

In my above example, for instance the work out by recompressing the residual steam to inlet pressure and temperature and then heating it to 1100 degrees is 21% more than compressing it to inlet pressure and 1100 degrees.

This would be adding 53 BTUs of heat to get 38 extra BTUs of work back a little 71% efficient move on 1/3 of the steam. Might deserve further investigation.

Good day ------ Bill G.

Bill, Harry,
What also needs investigation is the use of different timing for the HP and LP valve events. None of the old car builders seem to have noticed this, except for one, the Lamplough-Albany in England in 1903.

I bring this up again, because long ago Besler did just that with the second airplane engine, a DA piston valved compound. He did it just as an experiment for for one good reason. His theory was that one should see the two PV diagrams perfectly match for maximum USABLE expansion in a given engine at all cutoff points and all load conditions. He was of the opinion that with fixed cutoff points on the valve gear, the two cylinders were not going to give the optimum expansion, vs. separate timing.

The engine had several piston valves for the LP cylinder, each with a different cutoff point for each of the valve gear cutoff positions, the HP remained fixed.
It was equipped with pressure transducers (Ditzler) at each end and we watched on a dual channel scope.
When the curves met each other, the water rate was lower than when they did not match.
The conclusion was that indeed matching the expansion curves did improve the water rate and it also meant separate valve gear for each cylinder to properly achieve this.
Changing the valve's cutoff point was not quite optimum compared to separate valve gear; but it did prove his point in doing this experiment.

Unfortunately, when Ann Besler gave her father's papers to the Bancroft Library, the person who "evaluated" the documents, highgraded many of them, including the test results from this experiment. I looked long and hard for them, they are missing now.
JC



Edited 1 time(s). Last edit at 12/20/2005 09:32AM by James D. Crank.

Hi Jim

You are right on. The stages need seperate valve timming control to be efficient over any significiant power band. Compounds can be designed the old way if the power output is expected to fairly constant.

The main problem is avoiding over expansion. I don't have anything that figures backword flow (flow into the cylander during exhaust) But the efficiency goes to crap when the end of expansion is below exhaust pressure. The pumping of exhaust could only make it worse.

Andy

Hi Bill

My non-compression cycle there still figures a mix of inlet steam with the residual. The clearance has two components to it's loss. The volume that must be filed and the low temperature,pressure residual steam that it mixes with. My cycle calculator shows the temperature drop(non-compression cycle) or gain(compression cycle) of that mix at the cutoff point.

Andy

Jim,

The 2 stage Williams in a way does this. There are no 2nd stage inlet valves and absolute minimum transfer port (reciever) volume. The second stage then also recompresses to the exhaust pressure of the first stage.

The net effect is the same as a single stage Williams with an extremely large usable expansion ratio. The PV curves mesh as a single contiuous expansion with only a change midway in the volume of the compression curve, the pressure constant at that point. This because the 2nd stage has to recompress more steam than the 1st stage does. About 1/3 residual for the 2nd and 1/10 for the first.

I imagine that by matching the PV curves in the compound that the least amount of free expansion is effected, thus increasing efficiency.

Andy, I don't know the difference but my 2nd stage (final stage) efficiency doesn't go to crap from over expansion. Unless of course there is a backflow of exhaust into the cylinder (deadly). This is stopped with reed valves (if it proves neccessary). If there is any over expansion the piston simply recompresses to condenser pressure, the reed valves open, and pushes the steam out. The same as adjusting the expansion ratio.

At about 2/3 exhaust the valves close for recompression.

Howard, Thank You, for whatever cosmic reason the close details on the Skinner duel cam system havn't come my way.. The cam system geometry and layout are finished and I am now starting on real world details of loads and inertias. Once that is finished I can finalize the cams ramp profile. That will tell me the timeing of the thing.

I believe I have pretty much maximized the cam systems basic layout by keeping valve train weights to a minimum. (other than pnumatic springs) I hope to answer the question of what useable minimum cutoffs are available at what RPMs with a totally variable cam system.

Thanks Everybody ------------- Bill G.

Hi Guys,

Here are the Williams numbers.

operating pressure -1000psi
operating temperature - 1000 F inlet
compresses to 1500 F
When the 1500 F compressed steam mixes with the 1000 F inlet steam the result is 1300 F steam.
300 F exhaust temperature
6% cut-off

These are the numbers that Williams got in their tests. If you want to debate this there is really nothing more I can add as these are the William's tests.

Howard

Andy,
Backflow should never occur if the engine is working with a vacuum in the condenser.
Something that another Besler demonstration drove home most dramatically.
The same engine; but this time the LP diagram shot out far to the xero line and the water rate certainly went down. He did this demonstration for my benefit, one of the advantages of attending the Besler College of Engineering.

Without dyno testing an engine, with transducers in each end of the cylinders, the actual performance is only guessing. They reveal all that is going on.
JC

Hi Jim

I would think that backflow would occure when the end of expansion pressure is lower then the exhaust pressure. So in the case you describe you would not be over expanding thus no back flow. Simple phisics. Flow is from higher pressure to lower pressure. I didn't mean to imply otherwise.

The PV diagram I expect to see in a case of over expansion is the expansion curve going smothely down to the point when the exhaust opens it would then increase up to exhaust pressure as the pressure equalizes.

I don't think we have an argument here.

A vacuume is a way of eliminating over expansion. I don't think that a vacume can be maintained in an automobile application. I think it possable at light loads to have some vacuum.

Increassing condenser area may increase frontal area and thus drag. The efficiency gain could easly beloast to increased drag. After it is fuel consumption we are interested in. Increased efficiency at the expanse of higher arodynamic drag could be a loss.

It's tuning these types things that can bring steam efficiency into the 21 century.

How's your book comming along.

Andy




Edited 1 time(s). Last edit at 12/21/2005 10:03AM by Andy.

Andy,
A vacuum can certainly be maintained in a car system. It just takes a bit of power.
However, don't even think about Doble's lousy piston pumps, they didn't do the job, and not even the 4 cylinder one on E-14, that Becker got from the G.M. bus.
Only a small Roots blower, driven from the fan turbine works the way it should.
Up hill, lots of steam, fan turbine and vacuum pump going like crazy.

Book coming along splendidly, pictures selected, and now we have to select the drawings and reference catalogs to go into the appendix.
Printing supposedly around the first of May. Sold only on our own web site.

Merry Christmas and Happy New Year
Jim

Hi Bill, Don't worry about the Skinner patents. Tom said they are all expired.
So copy all you want. The company is out of business anyway.

Howard



--

Hello Howard,

I got a chance to review the or some of the Skinner patents. There are some major differences. Actually now I'm glad I didn't see them before I did my design. I've no cam roller that flaps back and forth like the Skinner. Can't imagine it doing high speeds.

I went to a distributor and got one of the Timkin catalogs. It has some cam rollers in it that I can start with. They only go up to a 1/2 inch width with a 3/4 inch diameter roller though and I was hoping to use a 3/4 width also. Although I believe that most automotive cam rollers are around 3/4 dia by 1/2 wide.

Who makes the little rocker rollers that push the valve stem? I found a company that makes cam rollers from silicon nitride which are a lot lighter than steel ones and might make a speed improvement by being lighter.

Merry Christmas ---------- Bill G.

Hi,,,Roller chain,,,1- 1/4 pitch,has a size,3/4x3/4,,maybee worth looking at,,,I'm still looking for 1-1/4p x5/8w x3/4d its an old size no longer made,,had hoped to find a usable length or old stock,lying in an old warehouse,,just dreaming again,,xxxx What has the Harley twin got for a cam follower, xxxx All the old quality cars used rollers,,not too hard to make,,We made a set for the old Panhard last year,,and fitted bronze bushes,,Cheers,,Ben

Hello Ben,

It looks like roller chain might be a good resourse for rollers, at least for prototyping. I found a rocker manufacturer using 1/4" amd 3/8" needle bearing cam rollers for the nose rollers. Can't find a manufacturer. I think the nose rollers (end of rocker roller) made of roller chain parts might work.

I found somewhere that titanium valves are normally coated with titanium carbide to make them wear and slide. Not on my radar yet.

There is far more to designing a simple rocker than I ever believed, everything has to be so light, and simple parts aren't common enough to be in the average bearing catalogs. Looked on the web a bit for your roller chain, most everything is standardized no luck.

Thanks -------------- Bill G.

Hi Jim

My understanding is: It takes more energy to pump a vacuume then can be gained in engine output. The basic laws of theodynamics make this so. You only get a percentage if the potential energy converted to work and inefficiencies of the pump means you can not convert all of that work back into pontential energy in the vacuume pulled. There would be a net loss. Stationary power plants pull vacuumes with large condensers.

I can see were water may be saved. Condensing more or all of the water would increae your water range.

It does take some energy to transport the exhaust through the condenser. More so in an automotive application as we can not relay on gravity. So how much more does it cost to pull a vacuume? Speculation.

If anyone can explain how using a mechanical means of pulling a vacuume would increase over all efficiency of an automobile. Please explaine how it world work.

Not trying to argue here. I just wont to know.

Andy

All files from this thread

File Name	File Size		Posted by	Date
ThrottlingTurnDown.pdf	58.1 KB	open | download	Andy	11/22/2005	Read message
ThrottlingTurnDown.pdf	62.3 KB	open | download	Andy	11/22/2005	Read message
hydro.jpg	123.5 KB	open | download	HLS	02/14/2006	Read message
P1010001aa.JPG	113.5 KB	open | download	Rolly	03/09/2006	Read message
P1010003aa.JPG	66.8 KB	open | download	Rolly	03/09/2006	Read message
P1010002aa.JPG	65.3 KB	open | download	Rolly	03/09/2006	Read message
SingleExpansion Vs Co.pdf	111.7 KB	open | download	Andy	03/10/2006	Read message
SingleExpansion Vs Co.pdf	111.6 KB	open | download	Andy	03/13/2006	Read message
HLS Vs Compound size.pdf	112.1 KB	open | download	Andy	03/13/2006	Read message
HLS Vs Compound =PD.pdf	111.7 KB	open | download	Andy	03/13/2006	Read message
HLS engine.pdf	64.8 KB	open | download	Andy	03/14/2006	Read message
HLS engine.pdf	73.6 KB	open | download	Andy	03/15/2006	Read message
Rankin.pdf	171.9 KB	open | download	Andy	03/15/2006	Read message
HLS engine.pdf	74.6 KB	open | download	Andy	03/15/2006	Read message
Endiing temp 850.pdf	28.2 KB	open | download	Andy	10/02/2006	Read message
Endiing temp X27.pdf	28.1 KB	open | download	Andy	10/02/2006	Read message
Fickett.JPG	66.9 KB	open | download	frustrated	10/05/2006	Read message
Over Expansion 1.pdf	24.2 KB	open | download	Andy	10/24/2006	Read message
FlowSpeed.pdf	23.7 KB	open | download	Andy	11/14/2006	Read message
Material.pdf	16.9 KB	open | download	Rolly	11/20/2006	Read message
white cliffs project engine.jpg	499.8 KB	open | download	grblake	06/30/2007	Read message
SV pickup.jpg	81 KB	open | download	Rolly	07/05/2007	Read message
112908ab.jpg	82.3 KB	open | download	Jeremy Holmes	11/29/2008	Read message
112908b1.jpg	87.6 KB	open | download	Jeremy Holmes	11/29/2008	Read message
Dieter engine.pdf	294.8 KB	open | download	Rolly	11/30/2008	Read message
Bryan Tractor.JPG	108 KB	open | download	Rolly	12/01/2008	Read message
Bryan Engine photos.jpg	84.6 KB	open | download	Rolly	12/01/2008	Read message
p1010002aa.jpg	36.4 KB	open | download	Rolly	12/02/2008	Read message
tractor1.jpg	136.6 KB	open | download	frustrated	12/02/2008	Read message
tractor2.jpg	111.8 KB	open | download	frustrated	12/02/2008	Read message
Tractor3.jpg	137.9 KB	open | download	frustrated	12/02/2008	Read message
tractor4.jpg	159.5 KB	open | download	frustrated	12/02/2008	Read message
tractor5.jpg	113.6 KB	open | download	frustrated	12/02/2008	Read message
tractor6.jpg	98.1 KB	open | download	frustrated	12/02/2008	Read message
071709a.jpg	77.4 KB	open | download	Jeremy Holmes	07/16/2009	Read message