Reciprocating Engine DetailsPlans of the Olympic Class Ships


Joe Burgard

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Oct 26, 2008
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First off, I'd like to say hello to all on this board since I just joined here. I find this site to be very informative.

After reading Samuel Halpern's article "Titanic Prime Mover", I thought I would like to take the analysis of the triple expansion engines further and contribute to my and for other people to understand the workings of these engines. I'm doing this from an actual design/mechanical engineer's perspective, so technical language will be used.

In order to do this I need specific details about the engines such as the clearance, piston clearance, indicator diagrams, etc, of each of the cylinders (Any of the Olympic class liners would do).

Where can I obtain this information and are there actual blueprints/plans of the engines?

I decided to do this since I found a mistake in his article about steam consumption (footnote 3) and I would like to further my understanding of triple expansion engines (will build one someday). There he calculates a rate of 6262 lbs/min of steam. However, if you take the energy available from 230 pisa (1200.8 BTU/lb) to 1 psia (1105.7 BTU/lb) steam for that mass flow rate, there is not enough energy to produce the 45,000 HP.

6,262 lbs/min * (1200.8-1105.7) BTU/lb *.023581 HP/(BTU/min) = 14043 HP

As you can see this is only 1/3 of the actual power output of Titanic's engines at full speed.
 

Bill West

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Dec 14, 2005
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Welcome aboard.
The turbine exhaust is about 25% wet so the energy left is only about 850 btu/lb. The start point needs to be extended roughly straight down to final pressure on a Mollier chart in order to find the quality of the exhaust steam that would fit the calculations. As the contact heater was rated at 700,000 lbs/hr and 6262lbs/min equals about one half of that, Sam’s flow rate is plausible, assuming lots of spare capacity.

I too have been interested in the clearance data for studying the engine performance but have had no luck in finding it. The known outline plans of the engines are:
-Engineering magazine, Jul-Nov 1910. The best period description we have, available in good libraries.
-The Shipbuilder, reprinted in the Ocean Liners of the Past books.
-the Hospital Ship Britannic site once had 3-4 detail drawings from a French journal of the era but the pages were dropped for some good reason a few years ago.

Bill
 

Joe Burgard

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Oct 26, 2008
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Thanks Bill.

I was just assuming that ideally the steam throughout the recip engines and turbines would stay saturated, keeping the condensation to a minimum since condensation will chew through turbine blades and can blow heads off cylinders.

With that much condensate adding heat to the other steam, a mass flow rate of 6262 lb/min does seem plausible.

Never really used a Mollier chart for steam (for air yes), always T-s diagrams. I haven't done a T-s diagram analysis, but obviously there will be some falling within the vapor dome.

Getting back to the clearance, I was curious since I was trying to draw a theoretical pressure diagram, then draw an actual one. See attached image.
134250.jpg


The reference books I have give general rules for clearance based on what type of engine, etc. Maybe I can reverse engineer the rough values based upon the required port passage ways, the types of slide valves, etc. We'll have to shoot ideas back and forth on this.

Other details you may know. What was the cutoff, release/exhaust point, compression, angle of advance for the cylinders?

In the article, Mr. Halpern gives the inlet pressures for each of the cylinders (230, 93, 39 psia for the HP, IP, and LP cylinders respectively) and an absolute back pressure of 9 psia. However for example, the pressure in the HP cylinder at release/exhaust has to be higher than 93 psia since there needs to be a pressure drop for flow to occur from the HP cylinder to the IP cylinder's reciever and due to frictional flow losses. I'm curious since it maybe possible to get a value of clearance knowing the exhaust and receiver(or valve chest) pressures.

Joe
 

Bill West

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Sorry Joe, bore/stroke/valve type/pressures/rpm/IHP is just about the entire extent of the information available on engine design. Even the cutoff is a guess, Sam thinks around 40% I think around 60%. I too tried drawing an indicator diagram about 8 years ago for studying fuel consumption but could not get beyond the typical values sort of diagram you’ve shown. There just aren’t enough details known.

Google Internet Archive, select texts and search for marine engineering and like terms. That should get you some period textbooks from which you can choose era typical values for interest’s sake. But after that it’s all guess engineering, I haven’t seen anyone prove out hard conclusions about the design details.

Bill
 
Mar 22, 2003
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Joe.

You asked about obtaining more information regarding the engines on the Olympic class ships. I would recommend the book 'Titanic The Ship Magnificent,' Ch. 15 on the propelling machinery. I believe that particular chapter was written by Scott Andrews who provide me with some of the detailed information that I used such as the pressures at the input to each cylinder in my on-line paper you referred to. Unfortunately, you will not find some of the details you are specifically looking for. I don't think they exist. And Bill may be correct in assuming a somewhat higher cutoff value than what I showed when running a full speed. Of course that will change the supply rate accordingly.
 

Joe Burgard

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Oct 26, 2008
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Bill, Sam, or anyone,

I was crunching some numbers in between studying for midterms and I would like your input.

First, while searching through period manuals I found that most HP cylinders of triple expansion engines have a cutoff of 60-75%.

Second is obviously a real indicator diagram (the dashed line in Fig 55a above) falls within the theoretical PV line (or saturation curve) meaning there will be condensation since there is not superheating. In "Elements of Heat-Power Engineering" page 212, they mention that from 20-50% of the steam admitted up to cutoff has been condensed. Is this fairly typical? I assume that in a triple expansion engine the condensation will be less since the temperature/pressure drop in each cylinder is less compared to single cylinder engines. In the diagram below, you can see that the quality at cutoff and end of stroke is about the same.
134289.jpg


Here's some calculations I did based on the assumptions of 70% cutoff and condensation of 25% at cutoff and throughout stroke (gives a quality of x=0.75).

Volume of HP cylinder = 99.402 ft^3
Volume at cutoff = 69.581 ft^3
Admission Pressure, P1 = 230 pisa
Specific volume, v_f1 = 0.01855 ft^3/lb
v_g1 = 2.000 ft^3/lb
Enthalpy, h_f1 = 368.4 Btu/lb
h_g1 = 1201.1 Btu/lb

Exhaust Pressure, P2 = 95 pisa
Specific volume, v_f2 = 0.017696 ft^3/lb
v_g1 = 4.654 ft^3/lb
Enthalpy, h_f2 = 294.45 Btu/lb
h_g2 = 1186.9 Btu/lb

Using the equations:
V_cutoff = m_vap * v_g1 + m_liq * v_f1

x = (m_vap)/(m_vap + m_liq)

yields:

m_vap = 34.683 lb
m_liq = 11.560 lb

While 11.56 lbs of water in the cylinder seems high (way too much to me) it only is equal to 354 in^3. And assuming a piston clearance of 9/16", the volume of piston clearance is 1290 in^3 so there is enough to avoid water hammering.

Using the enthalpy of the two states given the quality of 0.75 and mass flow rate of 46.243 lbs/stroke gives the energy per stroke of 29.4 Btu.

Converting to horsepower at 75rpm yields 4810 HP for the HP cylinder. Which is reasonable given a roughly equal partition of the total work (15,000 HP) among the cylinders.

Comments?

Problem with this is that unless the condensed steam is reheated, there isn't enough power to produce the remaining 10,000 HP given the mass flow rate.
 

Bill West

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You’re going into finer details than I have but indeed that initial condensation is countered on the next page by a re-evaporation resulting in little net change of quality.

With steam any expansion results in both temperature and pressure changes at the same time so we’ve got to use the Mollier chart to see what proportions apply our case. Expanding from 215psia to 95psia comes out as ~5% wet, ~1140btu/lb. This is what leaves enough heat for the following stages. Perhaps trying to space out a theoretical curve with ideal values first to get a broad fit would be better.

In Googling for a chart look out for those that are for air and those that are in metric. A British measure, steam chart in .tif format is available at http://www.conestogac.on.ca/~mwawzonek/thermoW2004/mollier/mollier.htm

The piston to cylinder head distance you mention might be fine but from a previous discussion involving the space needed between the valve and the cylinder for steam passages I’m coming up with a suspicion that total steam space clearance could be quite large, say 10% as a wild guess.

Bill
 

Joe Burgard

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Oct 26, 2008
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Thanks Bill for the Mollier chart, that one is much better (bigger and clearer) than one I already had.

I figured my next step will be to do a theoretical Pressure diagram, to see what the ideal values would be for each position and cylinder. The only problem I have is still the mass flow rate, it's too small. After looking at the Mollier chart, 5% condensation is not much (but that assumes isentropic expansion). My other thought is that while condensation due to wire drawing of the valve and cool cylinder walls happens, the great amount of latent heat released is a so much that it would stop any further condensation from happening.

To illustrate an ideal expansion of one lb of steam from 230 psia to 95 psia releases 14.2 Btu for work. At 230 psia a 5% percent condesation releases 41.64 Btu.

After the theoretical curve is done, I'll go talk to the head of the ME department (heads three thermodynamics journals) to see what he has to say.

Lastly in my last post the clearance I was refering to was the piston clearance: distance between the piston and the head at top dead center. From the various books I have, the clearance space is probably higher than 10% for the HP cylinder. In "Marine Engines and Boilers" it has that for piston valves for small cylinders (I assume "small" is relative here) with short straight ports a clearance of 12-18%. For small and medium cylinders with long and large ports a clearance of 18-30%. Also listed was for large LP cylinders with flat slide valves, a clearance of 8-10%.

More research to continue...

Thanks again for bouncing ideas back and forth.

Joe
 

Bill West

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Dec 14, 2005
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Half of Sam’s 6262lbs/min flow rate times your 55.84Btu drop is 4122hp for one HP cylinder, that is getting plausible. More so because I am reminded that the power could be divided by stages (3x5000hp) or by cylinders (4x3750hp). I want to check the Yarrow-Schlick-Tweedy description, I think it is likely the division is equal power per cylinder.

It will be interesting to see how much of reciprocating machinery your Department Head remembers!

Bill
 

Joe Burgard

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Oct 26, 2008
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Didn't think of that before...equal work in each cylinder instead of equal work per stage. Assumptions can bite you. Obviously equal work per cylinder would make sense with regard to getting a smoother running engine and even stresses on the crankshaft. I'll look into the Yarrow-Schlick-Tweedy system as well.

I'm curious too as to what the Dr. will say when I ask him about piston steam technology. He may just give me a weird look like why would you care about outdated technology. He must have received his degree in the late 1950's since he was about 10 years old at the end of WWII (I know that since he told me that his family fled from the Ukraine when it was sure the Soviets were going to reconquer the region in the face of the retreating German army). Obviously by the 1950's steam reciprocators were all but gone, so he may have not learned much about them.

Joe
 
Jun 11, 2006
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After reading this interesting discussion, I thought I would join in with my own thoughts and observations. I have no degree, only thirty years experience working with reciprocating steam engines for a living.

I agree that 42% is way too short a cutoff for full power operation of this plant, and that 60% to 65% was the more traditional cutoff used in triple expansion marine engines. As far as they go, Mr. Halpern’s calculations on steam consumption are correct. Since the HP cylinder at a specific cutoff contains a fixed volume, since a pound of steam at 230 PSIA occupies a specific volume, and since the only way for steam to get into the engines is through that HP cylinder, calculating the weight of steam used is cut and dried. Going with the assumption above regarding cutoff, we can assume that actual steam consumption may have reached 9400 lb. per minute. Since the plant is widely documented as producing 45,000 horsepower, there must be a way for that amount of steam to produce that much work. By the way, that amount of steam consumption at 45,000 HP comes out to 12.5 lb. of steam per horsepower per hour, in line with the documented consumption of engines of this type.

The way I see it, the engine only really starts running off of the heat in the steam after the point of cutoff when the steam starts to expand. Prior to that, the steam is simply the “hydraulic fluid” pushing the piston though nearly 2/3 of its stroke. In the IP cylinder’s admission stage, 60% of its volume is being filled by nearly 100% of the HP cylinder’s volume, not quite enough supply to meet the demand, accounting for “drop” in the IP receiver, but other than that, that steam is not expanding either, it is being pushed into the IP cylinder by the back side of the HP piston until the point of cut off of the IP valve at which time it starts to expand, converting heat into work. A similar situation occurs in the LP cylinder. Perhaps that phenomenon is the cause of the discrepancy between Mr. Bugard’s calculations and the historical facts.

When the engine has reached its maximum temperature after several hours of full power operation, each cylinder settles at a temperature somewhere between the temperature of the admitted steam and the temperature of the exhaust steam, minus any losses to outside radiation. As stated, within each cylinder, condensation during admission is counteracted by re-evaporation during expansion, meaning that each cylinder will exhaust dry steam into the receiver of the next stage. Accordingly, I believe that losses due to condensation of this very large engine when fully heated will be negligible. An interesting exercise in calculating them would be to calculate the BTU losses per square foot of the exterior cylinder and receiver surfaces per hour though their insulation and jacketing, and then calculate how many pounds of steam would need to be sacrificed to supply that loss.

Finally, yes, getting each of the four cylinders to produce equal horsepower would be a top priority once full away was rung up on the telegraphs. The engineers would proceed with taking indicator diagrams, doing calculations and linking in or out on each cylinder’s valve gear as needed to equalize the work done, probably every voyage. Smooth running, cool bearings and low fuel consumption depended on it.
 

Bill West

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Nice item from Google. Search “Google Patent”￾ then “Yarrow Schlick Tweedy”￾, up comes Mr Schlick’s patent and 2 more, in english to boot. Use the PDF download. Having found and read these with difficulty however they appear to only be balancing the machinery weights, not the cylinder power thrusts....

I’m left thinking the issue with the power thrusts is that they result in a twisting torque between the prop shaft and the engine frame. 8 pulses over each revolution torque then ease the frame, leaving a degree of vibration. Compared to the irregularities in propellor loading this might not be the biggest concern. Maybe all the issue here would be is one of evening out the average size/spacing of the pulses. With the present spacing it would involve medium changes to the horsepower distribution.

For the 106-100-54-100 degree crank spacing I get 4476.5 hp on the HP & IP and 3023.5 hp on each of the LPs. I had also roughed out an ideal energy distribution for the indicator cards and lo and behold the intermediate steam pressures that Sam got for his paper come very close to those horsepowers.

Time for some more thinking.

Bill
 
Mar 22, 2003
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Hi Kelly.

Thanks for your input. I tend to agree that I had used a cutoff that appears to be on the short side in my original calculations. The 42% was based on specific volume ratios of saturated steam at absolute pressures of 93 and 230 psia. But 93 was at the input to the IP, not the end of the HP stroke. A cutoff of 60-65% is probably a better estimate. I have seen cutoff values in that range for the operation of the battleship Texas running at high speed during her trials. The Texas had 4 cylinder triple expansion engines. With the higher cutoff, the steam consumption rate for the Olympic class ships would be about 50% greater than what I wrote in my on-line paper. (I'm trying to find the hyperlink to that data for the Texas, but what I have appears to be broken.)
 

Bill West

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I have an oops about cutoffs to explain today. I figured Titanic’s cutoff at around 60% and this agreed with typical period values. But I’ve been meaning to recheck my interpretation of how cutoff is expressed, sample figures are everywhere but a definition was rather hard to find. Is it “admission/60% of the stroke to go/expansion” or “admission/60% of the stroke gone/expansion”? Well of course it is the latter percentage but what I had erroneously used is the former. This means my opinion of 60% is misdescribed and should have been stated as 40%. This puts me in complete agreement with Sam’s figure.

The next catch however is why do we have 40% when period journals suggest 60% for marine work? They also use 60% for locomotives but stationary work of the era was 25%. My conclusions are that transportation couldn’t afford the space for the more efficient but larger 25% cutoff engines and just put up with the poor economy. In the Olympic/Titanic case however maybe the combination of higher speeds and a roomier hull made allowing for a physically larger engine a better deal than providing enough bunker space for the needs of an inefficient one.

So at this point, despite period marine texts we have a volume based calculation by Sam and a Btu based calculation by myself that both show 40% cutoff. Now we may both be knowingly overlooking the detailed inefficiencies on the indicator card but it does seem pretty clear that the Olympic class has to be a more efficient 40-50% cutoff rather than the era typical 60%.

Bill
PS Good morning Sam, I wrote before checking today’s posts but as you can see I’m not sure you need to ease off in your views very much.
 

Joe Burgard

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Oct 26, 2008
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Bill, there's actually two different kinds of cut-offs.

First is apparent cut-off which is the ratio between the portion of the stroke completed by the piston at the time the valve stops admitting steam to the cylinder to the total length of the stroke. For example: if the stroke is 20 inches, and the steam is shut off when the piston has completed 5 inches of the stroke, the apparent cut-off is 5/20=0.25 or 25%.

The second is real cut-off which is the ratio of volume of the steam in the cylinder when admission stops (apparent cut-off point) and the volume at the end of the stroke. Both volumes include the clearance volume of steam between the piston at top or bottom dead center and the valve. For example if the volume of steam in the cylinder including the clearance at the point of cut-off is 10 cubic ft, and volume including clearance at the point of exhaust is 20 cubic ft, then the real cut-off is 10/20 = 0.5. Real cut-off is used to calculate the ratio of expansion and is just the reciprocal of the real cut-off.

So the cut-offs we were refering to earlier (65% cut-off) are the apparent cut-off, while Sam's article was using the real cut-off (since he used the ratio of specific volumes).

It's enough to make your head spin, and it goes to show me how much I still need to learn here.

Kelly - Thanks for your input, your explanation of "hydraulic fluid" then expansion cleared up the issue. The HP cylinders basically have receivers of infinite volume (ie the boilers) to work from for the 2/3 admission then 1/3 expansion. Obviously the IP and LP do not have infinite volume receivers and adjusting the cut-offs in each cylinder then effects the admission and exhaust pressures to equalize the work. I'll be working out these figures to see if that gets roughly correct results.

To your point about calculating the heat loss of the engine at steady state (it's operating temperature after a long time of running) due to radiation, convection, and conduction is no fun task. REAL heat transfer calculations just suck to do. I agree that condensation should be minimal, just that my numbers weren't coming out right. But then again, you don't get the right answer by using the wrong equation.

Joe
 

Joe Burgard

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I was able to get some of those great drawings of Britannic's engines and I have been using them to get dimensions of the various parts. It takes a while to do.

I plan on doing a Diagram of Volumes for the engine. In order to do this I need the 4 inputs:
1. Cylinder Dimensions (Sam's article)
2. Positions of Crank (Sam's article)
3. Clearance and Receiver Volumes (trying to find from drawings)
4. Steam Distribution (will have to make an educated guess, like 66% cut off for the HP cylinder)

So far I've found the receiver volumes (roughly). The pipes were easy to dimension, but inside the valve chests and the valves themselves are much less accurate.

HP to IP cylinder R1=1.9*volume of HP cylinder
IP to LP cylinders R2=3.1*volume of IP cylinder

These values roughly correspond to the guidelines set in "Marine Engines and Boilers" on page 27. HP to IP receiver 1.8-3.8*volume of HP cylinder. IP to LP receiver 1.3-2.3*volume of IP cylinder.

Next to roughly figure the clearance volumes...

Joe
 

Bill West

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Here’s the contents of the data files for reference and some design comments:
Change the EditDimen/SetGraphicZoom to 150.00 if you are entering this. Just type over the stock file data and then save it under a different name.

HP -inside admission -file: BritannicHP1.dsi
......Frame Dimensions..................full size................................................full size
..................................................inches..................................................inches
.1...Driver...................Diameter..122.000....Rocker Upper..........Back Set.....0.000.....5.
......Stroke.....................Length....75.000....Rocker Lower.............Length..180.000.....6.
......Main Rod.................Length...164.000....Rocker Pivot.................Vert..179.950.....7.
......Eccentric Circle..............Dia....16.000.........."......"...................Horz..171.000.....8.
.2...Eccentric Dia..........................50.000....Valve Con Rod..........Length..130.000.....9.
.3...Eccentric Angle......ºForward....52.000....Lifting Link...............Length....81.000......
............"........"...........ºReverse....52.000....Reverse Arm Vert.....Length......1.950....10.
......Eccentric Rod.........Forward...171.000......,....".............".....Back Set....-0.102....10.
............."......."............Reverse...171.000....Lifting Arm..............Length...-40.000...11.
......Link.........................Radius..161.000....Rev/Lift Arm Pivot.......Vert......81.000.......
......Link Center Pivot...Back Set......0.000............"..........".............Horz...201.000.......
......Link-Ecc Rod Pivot........Vert....22.000....Rev Arm to Center......Norm......1.040....12.
..........."........."..........Back Set.......1.600.............".........."............Max......1.300....12.
. 4...Rocker Upper Vert...Length..-180.000...................................................................
-all the main dimensions are from the driver center.
-unworkable numbers lock up the data entry with an floating point error, you have to undo something.
-an "err" light also flashes at the top any time something doesn’t fit.

HP frame data notes:
1 -for looks only, the outside of the crank
2 -for looks only
3 -180º from locomotive position because of direct valve actuation. This is direct, inside admission so it is the "open rods" pattern (the rods are not crossed when the crank is pointing towards the cylinder).
4-made negative to fold this locomotive style rocker into itself. Effectively removes it from the mechanism as this is a direct actuation engine.
5 -leave at zero as part of folding the rocker into itself
6 -set high to make the valve con rod motion virtually horizontal
7 -trimmed to reduce blinking of valve con rod pixels
8 -center the rocker's swinging
9 -equals Cylinder ctr to Driver ctr less Eccentric Rod. This is 2" longer than actual because program can only center the valve on the main piston even though it is more to the left on actual engine. See more below.
10 -33"x1.724" ±36º arm cut short to fit. Rev Arm Cent position tips 3º to right.
11 -negative sets arm to right instead of left
12 - ± range fixed in program, not user adjustable.
200.029/201.069/202.109" is gear Fwd/Cent/Rev and gives 67.4/77.0% =72.2% avg cutoff in Fwd.
200.504" Fwd gives 43.2/56.8% =50.0% avg cutoff.

additional info for note 9:
The angles in the main rod to crank relationship are known to make the crank rotation and piston stroke uneven (sinusoidal). But less well known, the relationships are not symmetrical to each other either, because of the differing angles in the long and short triangles outlined by the 1/4 and 3/4 positions of the main rod stroke. This is called "angularity" and a consequence of it is that the cutoffs at each end of the cylinder are not equal. The Valve Con Rod Length choice has not tried to "re-center" this however because while it can vary the error, it can't eliminate it.

.......Cylinder Dimensions..............full size................................................full size
..................................................inches..................................................inches
.1...Cylinder ctr to Driver ctr........301.000....ValveBore ctr to ValvePort........24.000.......
.2...Cylinder CL above Driver...........0.000....ValveBore PortWidth.......Rear....5.000.......
......Cylinder Bore........Diameter....54.000............".........."...............Front....5.000.......
.3...Cylinder Bore...........Length....85.000.....Valve Ctr to Inside Edge...........................
......Cylinder Steam Port...Width......7.500............................of Inside Ring..20.500.......
......Piston ctr to Crosshead ctr....................Valve End Width (Outside..........................
............................(Piston Rod)..137.000............Edges of Rings)........Rear...9.000.......
.4...Piston.......................Width....10.000............"......"....................Front...9.000.......
......Valve CL to Cylinder CL............0.000....Valve ctr to Joint.....................................
......Valve Bore...........Diameter....35.000.............................(Valve Stem)....0.000....5.
.4...Valve Bore..............Length....74.000.................................................................

HP cylinder data notes:
1 -equals Main Rod + Piston Rod
2 -needs to be in line for direct actuation
3 -added piston width, not used in calculations
4 -for looks only
5 -for direct actuation this is all in the Valve Con Rod dimension in the Frame section

LP -outside admission -file: BritannicLP1.dso
......Frame Dimensions..................full size................................................full size
..................................................inches..................................................inches
......Driver...................Diameter..122.000....Rocker Upper Vert....Length..-180.000.......
......Stroke.....................Length....75.000....Rocker Upper..........Back Set.....0.000.......
......Main Rod.................Length...164.000....Rocker Lower............Length..180.000.......
......Eccentric Circle..............Dia....16.000....Rocker Pivot.................Vert..179.950.......
......Eccentric Dia..........................50.000.........."......"...................Horz..171.000.......
.1...Eccentric Angle......ºForward..232.000....Valve Con Rod..........Length..130.000......
............"........"...........ºReverse..232.000....Lifting Link...............Length...81.000.......
......Eccentric Rod.........Forward...171.000....Reverse Arm Vert.....Length......1.950.......
............."......."............Reverse...171.000......,....".............".....Back Set....-0.102......
......Link.........................Radius..161.000....Lifting Arm..............Length...-40.000.......
......Link Center Pivot...Back Set......0.000....Rev/Lift Arm Pivot.......Vert......81.000.......
......Link Center Pivot.........Vert......0.000............"..........".............Horz...201.000.......
......Link-Ecc Rod Pivot........Vert....22.000....Rev Arm to Center......Norm......1.040.......
..........."........."...........Back Set......1.600.............".........."............Max......1.300......

LP frame data notes:
1 -opposite of HP because of outside admission. This is direct, outside admission so it is the "crossed rods" pattern (the rods cross when the crank is pointing towards the cylinder).

.......Cylinder Dimensions..............full size................................................full size
..................................................inches..................................................inches
......Cylinder ctr to Driver ctr........301.000....Steam Port Width..........Rear.....7.000.......
......Cylinder CL above Driver...........0.000....Steam Port Width.........Front.....7.000.......
......Cylinder Bore........Diameter....97.000.....Valve Chamber..........Height...38.000....4.
......Cylinder Bore...........Length....76.500.....Valve Chamber..........Length...38.000....5.
......Cylinder Steam Port...Width......9.000.....Valve Height..........................19.000....6.
......Piston ctr to Crosshead ctr.....................Valve Length.........Rear Half....36.000....7.
............................(Piston Rod)..137.000.....Valve Length........Front Half....36.000......
......Piston.......................Width......1.500....Valve ExhCavity.....Rear Half....25.000......
.1...Valve Surface to Cyl CL...........64.000.....Valve ExhCavity....Front Half....25.000......
......Valve Stem CL to Cyl CL...........0.000....Valve ctr to Joint.....................................
.2...Exhaust Port............ Width......4.000.............................(Valve Stem)....0.000......
.3...Bridge Width..............Rear.....22.000.................................................................
......Bridge Width..............Front....22.000.................................................................

LP cylinder data notes:
1 -valve is left floating above the cylinder because the ports won’t stretch well enough. This is the actual dimension anyway, just picture it turned into line with the engine.
2 -arbitrary, for looks
3 -deducted 1/2 of exhaust port width
4 -diameter of valve piston
5 -for looks only
6 -shows half of valve piston
7 -to outside of rings

The valve is actually a piston type, the dimensions were adapted to the slide valve in the program as:
ValveBore ctr to Valve Port 24.000 -entered as Bridge Width less 1/2 of Exhaust Port
ValveBore PortWidth Front & Rear 7.000 -entered as Steam Port Width
Valve ctr to Inside Edge of Inside Ring 25.000 -entered as Valve ExhCavity Front & Rear
Valve End Width (Outside Edges of Rings) Front & Rear 11.000 -added to Valve ExhCavity size and entered as Valve Length

Optional drawing of lines for the actual Rev Arm:
ReferPoints needs to be turned on. This is troublesome to get entered, it wants to do it from reference points. So using them place cross1 and 2 anywhere with left and right clicks, then in DrawLinesCircles select a row, press the buttons to put in all but the location and then press Draw Close to connect it to the ReferPoints. Get anything you can in and then revise it to the correct numbers. There are buttons in the ReferPoints windows to remove them when you are done.
Data:
........No..Type..Color..Ref to...Line1Vert..Line1Horz....Line2Vert..Line2Horz
Fwd...1....line...green...0,0.......81.000....201.000.......109.943....185.148
Cent..2....line...green...0,0.......81.000....201.000.......113.955....202.724
Rev...3....line...green...0,0.......81.000....201.000.......107.985....219.996

Bill
 

Bill West

Member
Dec 14, 2005
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Although running the engine dimensions in the valve gear program suggests long cutoffs I note that the program also shows that the valve gear can readily be linked up to say 40-50%. A check of Brown's reversing engine shows that it's hydraulic cylinder will lock it in any position you want, just stop moving the control handle. So why design with 70-80% cutoff if you only need to link up to 40-50%? Well it would leave you room to easily adjust the power (up to the maximum that the boilers will support) should you have troubles in meeting the trials requirements or if future needs change. And at sea, if one engine fails it would allow forcing the other engine while you make port and would keep up the supply of exhaust to the turbine too. So I am satisfied that the long maximum cutoff need not exclude the possibility that design performance was based on a much shorter cutoff.

Having said that I have looked at typical water rates for period engines and think we are headed to 13 pounds/hour per IHP for a recip-turb combo exclusive of separate accessory pumps (Engineering Oct 30, 1914 pgs 538-39). For interest, other machinery types were no better than 15 lbs, except geared turbines but they were just starting to learn how to cut gears that could handle ship power levels. For the Britannic at 50,000hp 13 lbs comes out to 10,833lbs/min (650,000lbs/hr) and is in line with the condenser and other pump capacities given in Engineering. The surface feed heater capacity of 700,00lbs/hr may well just be the 620,000lbs/hr condenser air pump rating with the auxiliaries, generators (50,000lbs) and hotel load (16,080lbs) exhausts added in

My own doodling with ideal engine calculations is presently only at half of 10,333lbs. So when I allow for additional steam to cover the engine inefficiencies I am faced with the problem of getting it into the HP cylinder. Only a cutoff up around 70% would swallow that much steam and get it into circulation. So full stroke on the valve gear might be correct after all.

Bill
 

Joe Burgard

Member
Oct 26, 2008
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Bill,

The drawings of Britannic's engines you speak of I have as well. They're real good and I just printed them out on 8.5"x11", which gave them a rough scale of 1cm (elevation/top) and 2cm(side views) to 1m. I'll eventually get around to blowing them up. I agree, millimeters are much easier to measure drawings and scale from.

I just got in the mail a few days ago via Abebooks an actual copy of "Heat-Power Engineering" by Hirshfeld and Bernard (same edition as is online in pdf), so I'll be following their ways of doing PV diagrams.

In regards to steam consumption, my research as well all says 12.5-13 lbs/ihp-hour (and I'm sure we're looking at some of the same resources). Using that as a benchmark, the cut-off has to be at least 70%. Using 10833 lbs/min yields 36.1 lbs/stroke of HP cylinder. At 230 psia the specific volume is 2.00 ft^3/lb therefore cutoff is 72.7%. As mentioned before, I hope to get a good idea of the cut-offs for the IP and LP cylinders later with more calculations.

Thanks for the valve program and the other information. I'll have to go over that in detail later. I'm definitely going to use those animations since I'm building an Erector/Meccano Triple expansion engine with the Joy valve linkage, and yes I it will really run on steam (or compressed air).

Joe
 

Bill West

Member
Dec 14, 2005
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111
Joe,
Regarding having calculated the HP admission volume and cutoff, have you looked at the LP exhaust volume? 36.1 lbs/stroke HP intake is 18.1 lbs/stroke exiting each LP cylinder. At 9 psia that has a volume of 765 ft^3 in a 321ft^3 cylinder. Now it might be 58% wet which would reduce it’s volume enough to fit. But the energy drop for the engine then becomes 120,000 horsepower or the like. The engine is inefficient but not that inefficient. That’s too much energy to lose, something else is going on in our picture. How are we getting exhaust, that has not given up an excessive amount of heat to condensation, out of the LP without getting excessive volume? Perhaps the release pressure in the LP cylinder is around 20 psia and exhausting expands it to 9 psia so that it will be dry for the long exhaust pipes and the turbine?? More figuring....

For more reading, Internet Archive has an 1898/1908 book that seems to have a good discussion on the uncertainties of cylinder wall condensation. “Thermodynamics of the Steam Engine”￾, C.H. Peabody. Try pages 142, 199 and 202.

If you haven’t already researched it, here are the Olympic, Titanic and Britannic pages in Engineering that I know of:
vol 90 1910
-Jul 1 pg 14-15 Welin Davits
-Oct 21 pg 564-572 Olympic -hull & machinery
-Nov 10 pg 620-621 Olympic -hull & machinery
-Nov 18 pg 693(and facing plates!) -695, 698 Olympic -hull & machinery, boiler/engine/turbine/electric room plans
vol 91 1911
-May 26 pg 678-679 Olympic(despite article title/lead on Titanic) -general arrangement, deck plans
-Jun 2 pg 734 Titanic -launch announcement
-Jun 16 pg 789-792?? Olympic -accommodation
vol 93 1912
-Jun 14 pg 802-806 Titanic Inquiry -deck plans
-Jun 21 pg 847-850 Titanic Inquiry
-Jun 28 pg 884 Titanic Inquiry
vol 94 1912
-Jul 5 pg 3-12 Harland & Wolff -facilities
-Jul 12 pg 38-51 Harland & Wolff -facilities
-Aug 2 pg 154-157, 166 Belfast graving dock
-Aug 16 pg 237 Olympic -rebuttal letter on steering problem
-Aug 16 pg 239-241 Belfast graving dock
-Aug 23 pg 266-268 Safety Appliances at Sea Report -ice route map
vol 97 1914
-Feb 27 1914 pg 273-283 Britannic -hull & machinery, engine/turbine plans, deck plans, launch

Bill
 

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