Testing Models

I stated in my introduction post that I was interested in testing model steam turbines and engines. I use Onshape, a free (public files) online CAD system, for my design and Excel spread sheets for the analysis. My interest is in comparing the performance of the models with what is theoretically possible with proven engineering guidelines. I’ll start this thread with the testing of my model steam turbine which can be seen in my Album. This turbine is a tangential flow impulse turbine with a 7/8” diameter rotor. The rotor has pockets instead of blades that simplifies the machining. The turbine originally had Rulon bushings which I recently changed to ball bearings. My early testing was with steam and now I am using an airbrush compressor that makes the testing easier.

The first step is to determine the output of the boiler or air compressor. Since my boiler was designed for a running pressure of approximately 25 psig, I chose a sonic nozzle for my turbine. For this pressure and slightly superheated steam, the spouting velocity is only a small amount over the sonic velocity. I made a nozzle that could be inserted into a short tube connected to the outlet of the boiler. The bore of the nozzle started with the smallest drill I had, and I gradually increased the diameter until the relief valve stayed closed almost the full time. The final nozzle throat diameter was 0.028” diameter. I connected the boiler to the nozzle directly with no valves and shortest line, so that all the flow went to the nozzle with a minimum of pressure drop. Next, I completely emptied the boiler and put a measured amount of water in and timed how long it took for boiler to run dry. I repeated this several times to get an average time. I found that my boiler could produce approximately 1.4 lbm/hr of slightly superheated steam at a pressure of 25 psig. The steam exiting the nozzle was almost clear so almost no moisture. I found from the steam charts the amount of superheat needed to have dry steam at the outlet of the nozzle. This gave me an inlet temperature of approximately 380 F. I then checked for this temperature, pressure, and nozzle bore size what the theoretical flow is. The theoretical flow for these conditions is 1.35 lbm/hr. I used these parameters in the design of my turbine. When I decided to get an airbrush compressor for my testing, I connected the compressor to my turbine to find the maximum pressure I could run with all the air going through the turbine nozzle. I found the maximum pressure was 24 psig and the turbine temperature was approximately 70 F. For this pressure, temperature, and .028” nozzle throat diameter the theoretical mass flow is approximately 1.74 lbm/hr. These are my assumed maximum outputs for my model boiler and airbrush compressor.

The next step was to find a way to measure the power. I decided to put a propeller on my turbine and measure how fast it would turn and then see If I could determine the power. My first testing with steam and the Rulon bushings had a maximum speed of 14,500 RPM. For the EP2508 propeller I was using, an online calculator showed a power of 1.2 watts at that speed. This matched with the amount of power used by an electric motor and the motor efficiency given for this speed from another source.

After running the turbine on steam for many hours, I did a test with the airbrush compressor and the same propeller. The maximum speed was only 8, 580 RPM and the turbine stopped and restarted on its own a few times. I took the turbine apart and could see the shaft was scored and rough in the areas of the bushings. I decided to replace the bushings with ball bearings and made the necessary changes to my turbine. The test with the airbrush compressor, the same propeller, and the ball bearings yielded a maximum speed of 17,012 RPM. This corresponds with a maximum power of 1.92 watts and illustrates how important reducing friction is in these models.

I will add the testing of a Stuart Turner ST oscillating cylinder steam engine I have started in my next post.

The third step is analyzing the results. I wanted to know how the efficiency of the rotor pockets compared with the efficiency of the blades in a Terry turbine operating under the same conditions. I designed a Terry turbine rotor that meets all the guidelines for optimizing a Terry turbine given in a study of low output power turbines done by Dr. O. E. Balje for the Department of the Navy. This rotor is designed to run in the same housing, using the same propeller, and my airbrush compressor. The required keyway cutter needed to form the blades is expensive but available. Dr. Balje’s study gives methods for estimating the performance that I have found to be very accurate when compared with the results on a Terry turbine that was built and tested. The shop that tested the turbine had the equipment to accurately measure the power and steam consumption. Using the Dr. Balje’s guidelines, I can calculate the approximate efficiency of the existing rotor pockets based on the power that was produced and the power that was available. The measure of efficiency of a turbine blade is the velocity coefficient. The velocity coefficient is the ratio of the speed of the steam or air going out of the blades compared with the velocity entering the blades. The velocity coefficient for the existing rotor geometry, nozzle position and throat diameter, a power of 1.92 watts, an air flow of 1.74 lbm/hr, a inlet pressure of 24 psig, a inlet temperature of 70 F, and a speed of 17,012 rpm is 0.34. The velocity coefficient for the optimum Terry turbine rotor for these same conditions is 0.53. The overall efficiency of the existing turbine is approximately 10.8% and the estimated overall efficiency of the optimum Terry turbine is 12.5%. This test showed that the performance of the ideal Terry turbine blades was approximately 1.55 times better than the pockets of the existing rotor but the overall performance would be only 1.16 time better.

Great post Turbine Guy!

I'm working on a design for a scale model single stage steam turbine. The wheel will be a Curtis type, two row with reversing sectors. I could go either radial or axial split on the casing, but am leaning toward axial split as it will be much easier to assemble. I haven't finalized the wheel diameter yet, but it will likely be 2.5 to 3.0 inches.It's kind of governed by how small i can make the buckets and maintain a somewhat efficient profile. While it is a scale model and I'm not really concerned with overall efficiency, I would like to keep the water rate somewhat reasonable...

Looking forward to the next post in your series!

Tim

Welcome to the nuthouse TG

An alternative turbine system I ran into the other day

Tim, thanks for your kind remarks. You are taking on a much bigger project than my simple turbine. I had to keep my turbine simple due to my limited machining skills. The 17,000 rpm speed of my turbine is much lower than the speed required for maximum power even with the modest output of my boiler and airbrush compressor. If I understand you correctly, you plan on making a velocity staged turbine. This will improve the efficiency at lower speeds but the increased number of passes through the blades will make the efficiency of the blades important. The axial or radial flow impulse blades can be more efficient than the tangential flow blade, but both types require a large number of very small blades for good efficiency. Dr Balje in his guidelines put the turbines in three groups. Those with 96 blades were high efficiency, 48 blades were moderate efficiency, and 24 blades were low efficiency. Within these groups the required mass flow increased with increases in rotor diameter. My turbine has 24 blades and falls in the last category.

Ady1, thanks for your feedback. I looked at the Tesla type turbine but felt it was better suited to higher mass flow rates. I am interested in knowing if anyone has been able to get good results with low mass flow rates.

TG,

You understood it correctly - the design I'm working on is a Curtis type velocity-compounded style turbine, single wheel, with two rows of buckets and a stationary reversing sector between them. As you stated, the additional pass lets the machine develop more power at a slower run speed.

The trick is going to be how small I can make the buckets while maintaining reasonably accurate profile and velocity angles. It will probably mean going with fewer buckets on the wheel - I'm not overly concerned with efficiency, but I would like it to look as much as possible like the real thing.

I might start off with a single row wheel, as I can design the casing to accommodate either. That would be simpler and quicker.........hmmmmm, even with fewer blades, it might be interesting to compare efficiency of a single row to two row wheel.....have to think about that.

Tim

Tim,

I believe the first thing you need to decide is the amount of energy you will have \available to drive the turbine. If you are scaling the size and shape of the nozzles and blades, you can make a pretty good guess of the flow required. Do you know what you plan to power your turbine with?

Tim,

I believe the first thing you need to decide is the amount of energy you will have \available to drive the turbine. If you are scaling the size and shape of the nozzles and blades, you can make a pretty good guess of the flow required. Do you know what you plan to power your turbine with?

I have also ran several tests on my Stuart Turner ST oscillating cylinder steam engines. These were purchased in the 1970's and were bought as fully assembled ready to run and had not been disassembled prior to the start of the testing. I ran a very unscientific but revealing test to try to determine how effective the packing is for sealing pistons. I used my 7/16" bore and 7/16" stroke single cylinder oscillating engine for the test. I suspected that the leakage would be large enough that my small boiler would not reach full pressure with the rotation stopped and the steam port wide open. When the boiler started to make steam, I let the engine run to pass the initial condensation through. I then stopped the engine in a position that all the ports were blocked and adjusted the spring that pushes the cylinder against the standard. I had to almost fully tighten the spring to stop the leakage from the cylinder and standard interface. In this position the pressure quickly rose to the maximum pressure of 25 psig and the relief valve opened. I let the engine run again to make sure there wasn't any water build up. The engine started immediately and the exhaust was almost all dry steam. I then stopped the engine in a position that the inlet port to the top of the cylinder was fully open. In this position, the steam had only two places to leak. The first place was the cylinder to standard interface and it appeared almost no leakage was occurring there. The second place was by the piston packing to the exhaust side of the piston. With the engine stopped in this position the pressure wouldn't rise above 10 psig. I released the flywheel allowing the engine to run and then stopped it again in it's last position several times. Each time the engine immediately went to full speed and the exhaust was mostly dry steam. Each time I stopped the engine in this last position the pressure never rose above 10 psig. I expected the leakage would be large enough to pass all the steam the boiler was able to produce. What I didn't expect was that it could pass all the steam at such a low pressure. When I started this test, I expected it would take over 20 psig to pass all the steam. The reduction in pressure due to the leakage reduces the power significantly, but if you just want to watch the engine run without a load it doesn't matter. Also, I used the tiny pressure gage on the boiler that is very hard to get accurate pressure readings. My goal was only to get an estimate of how effective the packing is after it has been run a few hours. Based on this test, I am leaning toward O-rings for my next steam engine. I will take the cylinder apart and try to get a picture of what the packing looks like.

I'm intrigued as to how you know the exhaust was mostly dry steam? As I understand it measuring the dryness fraction of steam is not trivial, and normally involves a throttling calorimeter

Andrew

I added the picture of the packing to my photo album. Sorry about the poor quality. Photos are obviously not my strong point. Before removing the piston, I tried moving it back and forth with the ports closed and only felt a slight resistance. The packing appeared very smooth and completely intact. I didn't have any new packing, so I tried repacking the piston using its existing packing and some of the packing from my second ST engine. Before removing the piston, I ran the engine with a propeller attached so that I could record the speed and power as I did with the turbine. The propeller used on the ST is a APC 8x6 and the engine ran at a speed of approximately 1,500 rpm. The propeller required 0.4 watts to turn at the speed based on the propellers performance data sheets. This was very disappointing performance, since I expected much more power. I assumed the low power was due to the poor sealing of the piston packing. I packed the piston as tight as I could and tried moving it back and forth in the cylinder with the ports blocked and could barely move the piston. The resistance was much higher with the ports open than with the loose packing. I tried a run with the tight packing but the relief valve on my boiler started leaking. I decided to to all my testing with the airbrush compressor. I'll add the results of all the testing I have completed in following posts.

Andrew, I assumed the steam was almost completely dry because it was almost invisible coming out of the exhaust port. After it traveled a short distance I could start to see the moisture. When using steam, I ran the engine quite a long time before starting to measure the performance so that it would be at full temperature. Using propellers for a load source definitely didn't help. One of the most important reasons I had for doing the remaining testing with a airbrush compressor was eliminating this unknown.

TG,

I intend it to be able to run on either compressed air or D&S steam - not to exceed 50 psig inlet.

In a Curtis type Turbine, pressure is converted to velocity in the nozzles and remains relatively constant going through the rotor blades. The cross section of a nozzle looks somewhat like a venturi sliced on an angle - it tapers from the throat diameter to about 75% of the blade width at the discharge point. This is where the velocity angles come into play........the velocity angles of the nozzle and buckets need to line up correctly to get the best performance.

I'll try to post some diagrams when I figure out how to include them........

Tim

It was easier than I thought!

This pic is a cutaway of an actual Curtis type single stage turbine - this one happens to be a Coppus TF-9.. You can see the 2 row wheel, the reversing sector, and the cutaway nozzle. The hand valves on top are used to isolate one or more nozzles. Reducing the number of active nozzles to the minimum needed to maintain load insure that the pressure drop is happening across the nozzles, where is designed to, and not the governor valve.

Tim

Tim,

I would choose air based on my experience using both. My airbrush compressor puts out a very consistent air flow and starts with the push of a button. I'm able to use pre-lubricated ball bearings without worrying about the steam washing out the grease. I also don't have to worry about wet steam and rusting. I suggest you choose your power source and make a nozzle. You will need the maximum diameter of the nozzle for scaling the blades. You're right about using multiple nozzles and being able to isolate them individually. The supersonic nozzles are very sensitive to pressure drop and can become unstable with pressure ratios higher or lower than their design. For optimum performance the angle of the blades needs to take into account the speed difference between the stationary and moving parts. If you plan on only running at low speeds this won't be as important.

I set the airbrush compressor regulator to 30 psig and started the compressor while holding the propeller in a position where the inlet port was blocked and the pressure just verily held with the airbrush compressor running almost constantly. At 30 psig pressure almost all the air could leak out of the joint between the cylinder face and the standard with the tension spring almost fully compressed. When I released the propeller, the engine initially ran at a speed of 1,060 rpm and after a fairly long run the speed reached 1,100 rpm. The air pressure was about 20 psi for the entire run. Apparently, the airbrush compressor doesn't have enough mass flow to keep up with the leakage of the engine and reach the set pressure of 30 psig . Before I tried testing the engine, the tension spring was never near fully compressed so I must have run at relatively low pressures. So far, I have never reached the performance of the very loose packing (1,500 rpm) either with steam or air. The extra friction from packing the piston as tight as I could seems to lower the performance since the tension spring does not appear to be strong enough to run at a pressure above 20 psig. I wonder if the spring tension is not enough or if the contact surfaces are not parallel. I need to cure the leakage problem before I can make any useful tests.

I attached the hose from my airbrush compressor to my boiler outlet tube to check the relief valve. The relief valve is leaking even at very low pressures and opened at approximately 20 psig, not the 25 psig that I assumed, and is the maximum steam pressure my engines have run on. I set the airbrush compressor at 20 psig to match the steam pressure and tested the engine with the tight packing. I tried different settings of the spring that pushes the cylinder against the standard. The highest speed reached was 1,010 rpm with the spring almost fully compressed. When I stopped the engine and put the cylinder in a position that both ports were closed there wasn't much air leakage with the 20 psig pressure. The maximum speed was close to the maximum speed with the air pressure set at 25 psig (1,120 rpm) so increasing the set pressure was almost offset by the increase in leakage. The actual running pressure was almost identical for the 20 psig and 30 psig set pressures. Internal leakage, pressure drop through the ports, and friction have a large effect on this small engine. I ran another test to with the pressure set at 20 psig and got a maximum speed of 1,120 again.

I ran my oscillating engine with the tight packing off and on for about 2 hours. The maximum speed was 1,160 rpm. This is slightly above the 1,120 maximum speed of the last test. The bedding in and wear of the packing appears to have leveled off to a point I can consider this speed to be the maximum for my test of the tight packing. The corresponding propeller power for 1,160 rpm is 0.18 watts. I checked the external leakage with the piston in various positions and the engine stopped. With the piston in a position that all the ports were blocked the leakage was almost completely stopped and only a few oil bubbles were escaping.. When the piston was in a position where either inlet port was near it's maximum opening, the leakage increased considerably. In these positions the leakage through the mating surfaces of the standard and cylinder and the leakage by the piston and connecting rod took all the flow of the airbrush compressor at it's set pressure of 20 psig. I rotated the propeller with no air pressure and looked at the mating surfaces with a light behind them. I could not see any separation in all positions with no air pressure. I did the same thing with the air turned on and could see the surfaces separate when the inlet ports were near their fully open positions. Apparently the internal force pushing on the cylinder with air pressure is cocking the cylinder on it's support pin and causing the leakage. Stuart used a relatively loose fit on the cylinder support pin that makes it easier for this to happen. This leakage plays a big part in explaining the disappointing performance with the higher pressure.

I machined the piston of by second ST steam engine for a floating o-ring. I ran the engine with the floating o-ring off and on for about a hour and a half. The results were similar to the engine with the tight packing. After all the parts bedded in, there was almost no external leakage with the cylinder in either position where all the ports were closed. When the cylinder was in either position that the inlet port was near it's maximum opening the external leakage started due to the cylinder tilting on the standard. I have no way of knowing how the external leakage compared with the engine that has the tight packing but it seemed slightly less. The top speed was about 1,200 rpm right after the engine had been oiled and then settled down to a speed of about 1,150 rpm and would hold that speed. The steady speed of 1,150 was almost the same as the maximum speed of 1,160 rpm obtained by the engine with the tight packing that had run for quite a bit longer. When I changed the cylinder assemblies on the standard used for testing, I compared the force needed to move the pistons with the ports opened and closed. The force to move the piston with the ports open was about the same. The force to move the piston with the ports closed was much higher with the floating o-ring which appeared to completely seal the piston. The engine with the tight packing took a lot more force to move the piston with the port closed than with the port open so was still sealing very good. The external leakage caused by the cylinder tilting keeps me from getting a good test of any of the seals. I was pleased with the performance of the floating o-ring since it did better than I expected with the high leakage. I need to fix the cylinder tilting problem to get good test results.