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Turbine Guy29/05/2019 21:18:33
115 forum posts
51 photos

Hello Werner,

I thought about the statement you made about turbulent flow in nozzles causing complete chaos in the turbine housing. The only time I have got chaotic flow from the nozzle when operating on steam is when the boiler has carried over water. When a slug of water is carried over the flow is upset considerably and the turbine speed drops dramatically and will only get up to full speed when most of the liquid is cleared. I have to run water levels way below what is recommended by my boiler manufacturer to keep from getting water carried over when operating with maximum burner heat. Since the pressure you are using is higher than what I use and the boiler on your train is moving, I assume it would be even more difficult to keep water from carrying over.

The only other reason for chaotic flow I found is explained below. This was copied from one of my turbine books.

Sonic Nozzle Shock Waves

Turbine Guy30/05/2019 19:30:17
115 forum posts
51 photos

I revised my velocity staged turbine design from what was shown on the post I made 10/04/2019. While I was waiting on my new indexer I tried many combinations of changing the rotor diameter, pocket angle, nozzle angle, and rotor/nozzle position. The following drawing shows what I estimate to be my best combination.

Tangential Turbine 3 VS 2A

Turbine Guy30/05/2019 19:37:41
115 forum posts
51 photos

I finished the rotor as shown in the following picture.

Rotor 3 VS

Turbine Guy02/06/2019 19:04:16
115 forum posts
51 photos

I added the bores and nozzle hole in my turbine housing as shown in the following photo. We're up to the high summer temperatures, so there is only a few hours in the morning that my non air conditioned shop is comfortable. That's also when I do my yard work, play tennis, and fish so my machining is going relatively slowly. The nozzle hole is not in the correct position. I made a error in the machining and will need to try again. I can still use the housing and just add the nozzle in a different location. The turbine will have an unused hole but it shouldn't effect the performance if I get the next nozzle hole right. I must have the nozzle hole in the correct position to get any meaningful results.

Housing 3 VS 2

Werner Jeggli03/06/2019 23:52:18
21 forum posts
5 photos

Hello Turbine Guy,

Your turbine parts look beautiful, are supported by theory and are also well machined. I really hope it will pan out in the end.

I just returned from a G1MRA event in Toulon/France where I ran the PRR-S2 with 5 coaches for half an hour and everybody clapped their hands. I took along my multipurpose model steam turbine project and discussed it with anybody showing their interest (quite a few).

My ambition is to create versatile steam turbine, which could be used for generating purposes using two 0.8mm nozzles, mounted either at the top or the bottom of the turbine casing for turbo generator application. By using the second slot for a reverse nozzle, working the forward blades from the wrong side (awful engineering - but it works at the price of awful efficiency) it can also be used for a direct mechanical drive.

I'm quite confident that I can reach the peak 10 Watt shaft power target and will run the corresponding tests to prove it.

Below is a bad picture of nozzle inset drawing and a Toulon picture. Behind the flower you see the PRR-S2.

duese 1 links.jpg

toulon 2019-1.jpg

Turbine Guy04/06/2019 20:21:21
115 forum posts
51 photos

Hello Werner,
Thanks for the kind remarks, my model turbines are intended to be easy to machine but functional. I’m attempting to see if the methods I have used for full size turbines can be used for models. I’m also trying to find ways to get reasonable efficiencies with designs that don’t require extremely thin blades or costly special tools. An example I gave in my 18/04/2019 post showed what Dr. Balje’s guidelines suggest for an axial impulse turbine with the mass flow, available energy, rotor diameter, and speed of my last turbine test with my airbrush compressor. As you have demonstrated, machining the impulse blades is possible but far from easy. The end mill diameter required for this example is 0.013 in. (.33mm). Surprisingly, this diameter is available from McMaster-Carr as shown below. I have trouble not breaking the 0.028 in. (0.71mm) diameter drill I use for nozzles, so I doubt I would have anywhere near the skill required to make the recommended blades of the example.

Carbide Two Flute End Mill
Square-End, Uncoated, 0.013" Mill Diameter, 0.039" Long Cut
In stock
$31.44 Each

Using the turbine to turn a high speed generator makes a lot of sense. You really need very high speeds to get reasonable efficiencies. Experimental Flash Steam by J. H. Benson and A. A. Raymond described a turbine made by Jim Bamford that had a 3-1/4 in. diameter rotor that produced 0.47 BHP at 60,000 rpm with 9 oz./min water evaporation. This was tested on his water brake that would have required a speed much below the turbine speed so was after the gearing and probably included the pump power loss. The flash boilers require very high pump pressure and his boiler and turbine/gearing fit in a racing boat about 3 ft. long. There wasn’t enough information to calculate the efficiency. The steam engines are often compared by pounds of water per horsepower hour they require. Jim Bamford’s turbine used approximately 72 lb/HP-hr of steam. For comparison two 5 HP simple steam engines a Semple and a Reliable have steam rates of 49 lb/HP-hr and 40 lb/HP-hr respectively. Both used saturated steam.
Your nozzle insert looks very good. It has a smooth transition from the inlet to the throat and appears to have a good admission length.
I am very interested in getting the results from your tests and based on what you have already accomplished think you will meet your goal.

Turbine Guy05/06/2019 21:04:53
115 forum posts
51 photos

Jim Bamford’s turbine described in the last post was a De Lavel axial impulse design. No pictures or drawings of that turbine were shown in the Experimental Flash Steam book. The following drawing of the Stumpf type of tangential impulse turbine that was also designed by Jim Bamford was shown. This turbine has a 3.0 in. diameter rotor and  produced approximately 1.2 BHP at 95,000 rpm using 1.5 lb/min of steam. The steam rate for this turbine is 75 lb/HP-hr. Jim apparently realized he needed to run the turbine at the highest speed possible to get the most power. No reason was given for changing to the Stumpf type of turbine from the De Lavel type. Both of these turbines are outstanding in their performance. Since these turbines were used for racing, maximum power was the primary consideration. Bamford Stumph Turbine

Edited By Turbine Guy on 05/06/2019 21:13:07

Werner Jeggli07/06/2019 21:25:44
21 forum posts
5 photos

Gentlemen,

Today, I test ran the turbine and with 2 nozzles 0.8mm dia, 3 Bar boiler pressure. I got 41'000 rpm, 22V, 400mA DC output. Now I need also to determine (measure) the other relevant parameters over a 5 minute period like liquid gas consumption, steam temperature, -pressure and -throughput. I would also like to measure the steam temperature prior to entering the nozzles with a thermocouple in direct contact with the steam. However, the steam rushes by very fast, hence its pressure will be substantially lower than the boiler pressure. Does this make the steam more saturated (because of the lower pressure) and how can I take this into account?

The exhaust steam will be condensed in a water cooled spiral tube and then the condensate measured.

Any suggestions ?

Turbine Guy08/06/2019 01:25:40
115 forum posts
51 photos

Hi Werner,

I look forward to seeing all your data. I have a few questions about the information you have already given.

Is the 3 bar gage pressure (pressure above atmospheric)?
Is one of your two nozzles being used for reverse rotation or are both being used for the forward rotation?
Do you know the approximate efficiency of the generator and rectifier?
Is the generator turning the same speed as the turbine?

For saturated steam the temperature will be the saturation temperature of the pressure where the measurement is being taken. If you can read the pressure very close to the nozzle, you can get the saturation temperature for that pressure. The percent of moisture will increase as the pressure decreases. If you start with saturated steam at 3.5 bar gage pressure and expand it to atmospheric pressure (0 gage pressure) the moisture in the steam will increase to about 8%. The temperature of the steam at the 3.5 bar pressure is 148 C and at 0.0 bar is 100 C. This is for isentropic expansion.

Since you already got a power output of 8.8 watts from the generator/rectifier, you may have already achieved your goal of 10 watts for the turbine power.

Turbine Guy11/06/2019 21:25:47
115 forum posts
51 photos

I finished most of the machining on my new turbine housing as shown in the picture below. I only need to add a tube to the nozzle opening and drill two holes in a mounting plate to run the turbine. I designed this turbine to be able to run the first row of blades only by leaving the reversing chamber off. The air or steam can exit the first row of blades out the opening required for the reversing chamber. I want to do this so that I can see the performance of the rotor with the larger diameter and greater number of pockets. The testing I have done on the last turbine indicates the open pocket design has about the same performance of a Terry turbine with the same rotor diameter, number of blades, nozzle admission length, and available energy. My first rotor with 24 pockets appeared to match the performance given by Dr. Balje’s diagram for Terry turbines with around 25 blades. My second rotor with 48 blades also matched the performance for Terry turbines with approximately 45 blades shown in the diagram in the post of 14/03/2019. The new turbine rotor with 60 pockets has more pockets than the curves for the Terry turbine in that diagram. Since the number of blades or pockets has a very large effect on the efficiency, I expect my new turbine running with just one row of blades to exceed the performance given for the Terry turbines in the diagram of the 14/03/2019 post. Whatever happens, I want to see the performance with or without the velocity staging.

Housing 3 VS 9

Mike Tilby12/06/2019 07:18:48
avatar
14 forum posts
8 photos

Hi, can I chip in on this very interesting thread which has some impressive photos.

In regard to the design of nozzles. Although, as has been said already, convergent-divergent nozzles are theoretically required to get maximum velocity from steam expanding to below the critical pressure, there were two reports in Model Engineer where people had assessed steam velocity using an impulse plate where the jet was directed against a plate attached to electronic weighing scales. The first report was by Mr Southworth in relation to designing the turbine for his 5" g loco "Turbo" (Southworth (2000) M.E. 185 (4136): 638 - 640). He experimented with many different shapes of nozzle and concluded that the best result was with a plain convergent nozzle which out-performed convergent-divergent nozzles of various angles and lengths - all with the same throat diameter. His article prompted a letter from Tom Jones and Professor Bill Hall (Jones & Hall (2001) Model Engineer 186 (4138): 65). They reported coming to exactly the same conclusion in their own tests on injector nozzles. It seems possible that in miniature nozzles effects such as increased friction in very small diameter ducts make it difficult for the steam to attain supersonic velocities and if sub-sonic steam flows into the divergent part of the nozzle then it will be slowed down instead of accelerated. In both those studies steam from conventional loco boilers was used. However, for their turbines, Mr Bamford and Prof Chaddock used steam at very high pressures and temperatures from flash boilers, so I guess the convergent-divergent nozzles might have worked O.K. under those circumstances. So it seems to me that the question of nozzle design is very much open to doubt and, of course, Werner has had good results with plain nozzles.

Regarding small diameter end mills, I buy them on Ebay where you can often find packs of 10 0.5mm cutters for less than £10.

Regards, Mike

Michael Gilligan12/06/2019 07:39:29
avatar
13788 forum posts
599 photos
Posted by Mike Tilby on 12/06/2019 07:18:48:

In regard to the design of nozzles. Although, as has been said already, convergent-divergent nozzles are theoretically required to get maximum velocity from steam expanding to below the critical pressure, there were two reports in Model Engineer where people had assessed steam velocity using an impulse plate where the jet was directed against a plate attached to electronic weighing scales. The first report was by Mr Southworth in relation to designing the turbine for his 5" g loco "Turbo" (Southworth (2000) M.E. 185 (4136): 638 - 640). He experimented with many different shapes of nozzle and concluded that the best result was with a plain convergent nozzle which out-performed convergent-divergent nozzles of various angles and lengths - all with the same throat diameter. His article prompted a letter from Tom Jones and Professor Bill Hall (Jones & Hall (2001) Model Engineer 186 (4138): 65). They reported coming to exactly the same conclusion in their own tests on injector nozzles. It seems possible that in miniature nozzles effects such as increased friction in very small diameter ducts make it difficult for the steam to attain supersonic velocities and if sub-sonic steam flows into the divergent part of the nozzle then it will be slowed down instead of accelerated. In both those studies steam from conventional loco boilers was used. However, for their turbines, Mr Bamford and Prof Chaddock used steam at very high pressures and temperatures from flash boilers, so I guess the convergent-divergent nozzles might have worked O.K. under those circumstances. So it seems to me that the question of nozzle design is very much open to doubt and, of course, Werner has had good results with plain nozzles.

.

Useful references, Mike yes

Aphorism : You can't scale nature

MichaelG.

Edited By Michael Gilligan on 12/06/2019 07:41:08

Turbine Guy12/06/2019 17:16:52
115 forum posts
51 photos

Mike Tilby,
Thanks for adding your valuable information on sonic nozzles. The section view of Jim Bamford’s turbine shown in the 05/06/2019 post shows a sonic nozzle. This drawing cautions that the nozzle position is not to scale. I think that caution was added to make clear the section through the nozzle had been rotated relative to the section of the rest of the turbine and the flow from the nozzle would be in the opposite direction relative to the rotor pockets. If he did use a sonic nozzle for this turbine it would definitely make the point that sonic nozzles can be used for pressures much higher than the sonic pressure. I have designed turbines with supersonic converging/diverging nozzles that worked well but the lowest power of these designs was about 8 hp (6 kw). These nozzles also had a straight section after the diverging section to stabilize the flow before it entered the inside of the turbine housings.

Mike Tilby13/06/2019 17:23:20
avatar
14 forum posts
8 photos
Posted by Werner Jeggli on 07/06/2019 21:25:44:

Gentlemen,

Today, I test ran the turbine and with 2 nozzles 0.8mm dia, 3 Bar boiler pressure. I got 41'000 rpm, 22V, 400mA DC output. Now I need also to determine (measure) the other relevant parameters over a 5 minute period like liquid gas consumption, steam temperature, -pressure and -throughput. I would also like to measure the steam temperature prior to entering the nozzles with a thermocouple in direct contact with the steam. However, the steam rushes by very fast, hence its pressure will be substantially lower than the boiler pressure. Does this make the steam more saturated (because of the lower pressure) and how can I take this into account?

The exhaust steam will be condensed in a water cooled spiral tube and then the condensate measured.

Any suggestions ?

Hello Werner, I would think that the drop in pressure at the entrance to your nozzles would result from friction along the pipe and from loss of heat by convection from the outside of the pipe. Of course both these effects will depend on diameter and length and nature of the surfaces of the pipe. Drop in pressure will mean a lower saturation temperature and friction will increase the heat content of the steam. So these effects will act to keep the steam dry. However, loss of heat by convection will tend to lower the superheat and/or increase wetness of the steam. It seems to me that the only way to be certain of the steam condition would be to ensure the steam is initially superheated to an extent that it remains superheated when it reaches the thermocouple at the nozzle entrance. If its temperature drops to the saturation temperature you will not know how wet the steam is since its temperature will not decrease further until it has essentially all condensed. I guess the only way to be sure of the pressure at the nozzle entrance is to have a pressure gauge attached at that point and then you will be able to look up the saturation temperature. (On my boiler I have a thermocouple at the entrance to the nozzle and I'm currently making some connections for pressure gauges before and after the first nozzle of the turbine).

No doubt I've over-looked something in the above waffle so I am ready to be corrected.

Mike

Turbine Guy14/06/2019 19:11:27
115 forum posts
51 photos

I ran my new turbine without the reversing chamber with my airbrush compressor and it obtained a speed of 21,500 rpm turning the EP2508 propeller. My estimation of the power required by this propeller to turn at that speed is 3.4 watts. The power available to the turbine from my airbrush compressor is approximately 18 watts so the efficiency is approximately 19%. This is an increase in power of about 1 watt over what was obtained by my last turbine for the same amount of energy. The increase in power is primarily due to the increase in rotor diameter, increase in number of pockets, and reduction of the pocket inlet angle. The increase in the rotor diameter was from 0.892 in. to 1.226 in. This increases the rotor tip speed and torque. The increase in rotor tip speed raises the power but also increases the rotational (windage) loss. The rotational loss is still low enough at 21,500 rpm that even with the extra set of pockets it did not appear to be excessive. The increase in number of pockets was from 48 to 60. With single nozzles, the number of pockets under the nozzle discharge opening becomes very important since energy is lost in filling the empty pockets. The pocket inlet angle decreased from 30 degrees to 25 degrees. The smaller the angle, the less energy is lost from the flow not being perpendicular to the direction of travel. The power achieved was greater than expected so my estimations of power from the propeller might not be conservative. However, the increase in performance with each of the changes using the same propeller and airbrush compressor verifies that Dr. Balje’s guidelines are useful in the design of model turbines.
The following photo shows my test setup. Not very pretty but it worked.

Tangential Turbine 3 VS Test

Werner Jeggli14/06/2019 20:40:17
21 forum posts
5 photos

Hello Mike,

I do not think the loss in pressure between boiler and turbine entrance is only due to friction losses in the piping. The main difference must be the more or less static steam pressure in the boiler and the reduction of pressure in the turbine inlet pipe due to its reduced cross section and therefore high speed of the medium (steam). (airplane wing effect). Pressure loss will however be compensated by an increase in entropy (the molecules move in the same direction). How to take this mathematically into account is above my capability.

Mike Tilby15/06/2019 16:31:24
avatar
14 forum posts
8 photos

Hi Werner

From my understanding, increase in entropy of the steam as it flows through the pipe will be due to friction and that will give a rise in thermal energy content of the steam which is basically random motion of the water molecules. In a perfectly efficient nozzle there is no increase in entropy as the steam accelerates and all the available thermal energy is converted to kinetic energy as the molecules tend to move more in the direction of bulk flow.

When you say “airplane wing effect” are you referring to loss of pressure in air as it accelerates as it passes over the upper surface of a wing? If steam pressure drops as it flows along the pipe then the steam will have a larger specific volume (i.e. volume occupied per kg). Since mass flow rate is constant that means the volume flow rate must increase and so the velocity must increase. That will require energy and so the temperature of the steam will decrease. A small acceleration of low velocity steam will require very little energy since kinetic energy is propotional to velocity squared.

If I understand you correctly, then you think the velocity increases quite a lot in your pipe. Does that mean the pipe is very narrow?

But whatever loss occurs in the pipe, is it not the steam condition at the nozzle entrance that is important? As Turbine-guy said, if you attach a pressure gauge just before the nozzle, then you will know the pressure and temperature at the point where it matters.

What is the internal diameter of your supply pipe? Is it very small? Knowing the temperature and pressure at the nozzle entrance would allow you to calculate the velocity in the pipe and hence the kinetic energy in the steam. However, it is usual for any pressure drop along the main supply pipe to be small because the pipes are generally much larger in cross-sectional area than the narrow point of the nozzle. By the time steam reaches the narrow point of the nozzle it has expanded to a lower pressure and so the volume flow rate will have increased a lot. This and the small size of the nozzle means a high steam velocity. In the supply pipe the pressure is high so the volume flow rate is small. This fact plus the large cross-section of the pipe means the velocity is low.

I'll be interested to hear if you agree with any of this.

Regards, Mike

Turbine Guy16/06/2019 20:06:36
115 forum posts
51 photos

I also ran my new turbine on saturated steam at a gage pressure of 50psi (3.4 bar) and an estimated water rate of 0.55 oz./min. The maximum speed it turned the EP2508 propeller was 28,000 rpm. The estimated power required for the EP2508 propeller at this speed is approximately 9 watts.

Ian S C17/06/2019 16:10:42
avatar
7438 forum posts
230 photos

A generator run on a small turbine, or any small power source would be pushing it to get to 50% efficiency, so there will be quite a difference between the electrical output and the brake HP measured with a Prony Brake, an interesting little experiment on it's own.

Ian S C

Werner Jeggli18/06/2019 16:59:16
21 forum posts
5 photos

Hello Mike,

I really do not feel at home in the pressure/enthalpy/entropy world. I'm trying to follow Norman Billingham's excellent series of articles in the SMEE Journal, but it's hard work and my grey cells are old and probably encrusted. If you tell me the measurements at the turbine entry can be taken at face value for further calculations - that's fine.

Yesterday, with the help of 2 friends (to also take simultaneous readings) we ran the test with 2 nozzles in action - and it was a total failure. The boiler couldn't keep up with the 2 nozzle steam demand. We will have to repeat it. With just one nozzle it should work.

To give you an idea of the hectic involved - have a look at the set-up!

testsetup 09.jpg

and the expected data

testsheet 01.jpg

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