Turbine exhaust steam quality 1

[tgmcg]

I'm looking for suggestions regarding the minimum recommended exhaust steam quality (for a condensiing steam turbine..obviously) consistent with a 3 year uninterrupted run without experiencing significant loss of turbine efficiency. The objective is to limit blade erosion caused by condensed water droplets. This is clearly affected by blade metallurgy and coatings (eg. TiN) used in the low pressure stages. While my gut instinct tells me that 93-95% steam quality might be a reasonable target to shoot for, I'd prefer to take a more scientific approach and seek input from other specialists in the field.

We're considering specifying a minimum allowable design steam quality as a reliability & efficiency criteria, and also to help compare apples-to-apples between competing vendor bids.

This criteria is intended to apply across condensing steam turbines of all sizes, speeds, steam conditions, steam path designs and makes.

It has been my experience that vendors sometimes attempt to use a lower (and unstated) exhaust steam quality as covert means to exaggerate their turbines power output & efficiency.

Any and all comments are welcome.

Regards,

Tom

 

[rmw]

Qualilty happens. The laws of Physics determine it.

You should use a Mollier Diagram to plot out your turbine expansion line to the design exhaust pressure. Remember that there is a slight "J" hook on the end of the line, but ignore that factor for now.

That will show you two things.

First, it will show you where you are in the moisture region with respect to expected quality for your design back pressure.

Second, it should make it obvious that in order to have less moisture, you have to raise your condensing pressure.

You mention turbine efficiency. Having a higher back pressure is counterintuitive to maximizing efficiency. But, it will minimize the moisture that you have in the back end of the turbine, if that is really your goal.

It also reduces the available energy from your steam. This is not something that I find my clients typically wanting to do. Usually, just the opposite.

There is a penalty for moisture based on the formula;

Delta E.L.E.P. = Exhaust loss X 0.86 (1-Y) where;

E.L.E.P. is Expansion line end point
Y is percent moisture at expansion line end point.

This formula modifies a value from a curve of exhaust loss that the turbine manufacturer will have to give you. That curve is based on annulus velocity.

That curve is derived from;

Annulus Velocity = Q X V X (1-y) / 3600 X A where

Q is condenser flow in lb/hr.
V is saturated dry specific volume in cu ft/lb corresponding to the actual exhaust pressure. (if end point is in superheat region, V is actual specific volume.)
A is annulus area in ft^2.
Y is percent moisture at ELEP as above.

Bottom line, the moisture at your ELEP is determined by the laws of thermodynamics, not the turbine manufacturer.

The best you can do is to make them design their equipment as best that they can (metallurgy, coatings, wear strips, etc.) to accommodate the moisture that Mother Nature is going to give you, or lower your efficiencies and available work accordingly with your cooling water flow or condenser sizing.

The condenser design does effect your condensing pressure, but again, that is not something that your turbine vendor has control of. They have to work with what they are given.

rmw

 

[tgmcg]

rmw,

Thank you for the helpful suggestions.

Using our SteamPlant software we can easily calculate the bulk exhaust steam quality with the accuracy of the IFC equation of state for water for any set of turbine operating conditions and isentropic efficiencies. Resulting steam quality is one of the outputs from the SteamTurbine model.

Exhaust steam quality is affected not only by condenser pressure, but also inlet superheat, inlet pressure and turbine efficiency...as well as turbine load (let's assume for now that the turbine is fully loaded). Turbine output power is strongly affected by exhaust steam quality. Pulling harder on the condenser results in higher power output and lower steam quality, and visa versa. However, as exhaust steam quality declines, the potential for blade damage in the lower pressure stages is also higher...it's probably a nonlinear relationship.

While different turbine designs may be more or less sensitive to droplet impingement damage in the lower pressure stages, it makes sense to attempt to limit this damage through the use of harder blade materials, flow path modifications, coatings and by placing reasonable restrictions on exhaust steam quality at the design stage. All things being equal, it would appear reasonable to expect that an exhaust steam quality of 85% will result in significantly more damage than an exhaust steam quality of 97%.

Not placing a lower limit on exhaust steam quality is in some ways analogous to not placing an upper limit on exhaust temperature from a gas turbine. Without placing such a limit, turbine life will be compromised in exchange for higher power output. In these days of increased focus on reliability, efficiency and extended run times, it makes sense to evaluate the technical and economic benefits of such a limit and to find an optimum value of exhaust steam quality for each turbine.

It seems such an obvious question that this wheel has probably been invented an re-invented many times over.

Best regards,

Tom

 

[rmw]

tgmcg,

Well, I have been cogitating on this one all day and have finally arrived at this. Are you addressing a specific problem or trying to have an academic discussion regarding whether or not a wheel has yet been invented.

First, let me amplify my earlier post. I recommended the archaic method of plotting a Mollier diagram because it shows the dilemma graphically. Not for the lack of software like SteamPlant or Steam Master/Pro or Turbine by Katmar et al.

A look at a Mollier diagram will show that where you are wanting to be is over close to the right hand margin under the saturation curve, and where you are trying to stay away from is low-mid page deep under the saturation curve.

Now lets stipulate that the less efficient turbines, single stage units and multi-stage (as in not many-stage) units will have (1) a beginning point in the mid pressure range and a superheat in a range much less than the practical maximums, and (2) a expansion line slope such that the end point will be just about where you want to be.

So small less efficient machines meet your criteria thusly. You won't get excessive moisture out of them at any realistic back pressure, generally speaking.

That leaves the large utility turbines, which I think are the real target of your quest.

Now in a theoretical discussion, the large steam turbine you would like to have in a perfect world would have a vertical expansion line (no change of entropy), but alas, we know that the perfect world does not exist.

That said, expansion lines for LSTG's approach vertical much more so than smaller turbines.

Now to comment on your analogy about the CT exhaust temperature limitation; true, that is a parameter that insures that the temperatures in the turbine sections do not exceed the temperature limitations of the metallurgy.

Steam has much the same thing. Practically speaking, 1050-1100 is an upper limit for super heat temperatures and 3650 psig is a practical upper limit for super critical units. Without going back and doing all the research, I know that these limits are being bumped up all the time, but only incrementally.

So, like the CT exhaust temp, Steamers have metallurgical limits.

That said, the beginning point for the expansion lines (for utility size turbines) are way to left and upward on the mollier diagram.

So, how do you get from there to where you want to be with a turbine having LSTG efficiencies?

One, you lower the efficiency, so that the slope of the expansion line will increase ending up in a region with less moisture, but since that results in lots less available energy and takes more specific energy, that is not a good idea.

Since these type of machines have an expansion line for the HP turbine typically and another one for the IP/LP turbines, we have to work with those.

The HP turbine ELEP is limited by the specific volume of the steam that has to be sent back to the re-heater along with the available energy considerations for the IP/LP turbine(s).

One thing that I can see is that a double re-heat machine will have 3 expansion lines, with each reheat path shifting the final expansion line over to the right so that its ELEP will be closer to the region you want to be in.

Another solution would be to do what the Nuc's do, and that is to take the wet steam and strip out the moisture and reheat it with live steam. Their fuel costs are such that they can get away with that. I doubt the economics are there to do it in a fossil unit.

Short of that, I think you are stuck having to deal with the moisture with the physical construction of the machine or with the condenser configuration.

As a person with turbine and heat transfer experience, another consideration that gives me pause regarding moisture in exhaust steam is the damage done to the condenser. I have seen some really banged up condensers due to moisture in the exhaust steam.

It is particularly bad on large machines that were designed for base load which now have to run load following and get down to some horrendously low loads at night. The wet steam, and low loads means that the wet steam leaves the turbine exhaust and dives straight down to the condenser tube bundle dead center (well maybe a little offset to account for the rotation of the exiting steam) and trashes the outer rows of tubing right under the exhaust hoods. Not a pretty picture.

Also, since most heat sinks, rivers, lakes, cooling towers, etc are seasonally affected, meaning that to be properly sized for summer operation when CW is hotter, winter conditions make for a really cold condenser bundle.

So, as I see it, the way to get where you want your criteria to be is to monkey with the condenser. Design your condenser/CW system so that the cooling water can be throttled in order to keep the back pressure up to where you want it to be in order to minimize turbine back end moisture.

Many times in the winter, and especially at night, one CW pump and one CW loop in a double flow condenser will be taken out of service in order to keep the condenser back pressure from getting too low. (The other side of the curve I referred to in my first post.)

You would have to come up with some way of modulating the condenser vacuum with CW flow so that when you need the turbine to be its most efficient in the summer time, it could, but when you didn't need it, it wouldn't have to be without your paying the moisture penalty.

That is just my take on it.

I did have one other thought. By specifying to a turbine manufacturer that your criteria was such and such minimum moisture content, that might be license for them to just lower their efficiencies, shifting the slope of the expansion line, and putting the ELEP farther to the right.

So, there you have my thoughts. But I enjoyed the mental exercise. It made 5 miles on the treadmill this afternoon go by real fast.

rmw

 

[tgmcg]

Hi rmw,

Thank you so much for your thoughtful post. I believe you see what I'm driving at.

My question is not academic, but rather an attempt to define some sensible operating limits for about 10 steam turbines I am currently re-rating for a refinery upgrade project. Average turbine rating is about 5MW @ 14,000 rpm...not large. The most effective means for optimizing unit reliability and efficiency is through sound engineering and design. That's my job.

We all know that droplet impingement is responsible for most LP stage damage, yet a consensus might not exist regarding how to limit this damage through the specification of favorable inlet and exhaust steam conditions. API 612 is silent on this matter, as are Bloch et al. It's usually up to the end-user to establish reliability and efficiency criteria. The turbine manufacturer has little incentive to recommend a lower limit on steam quality, as more turbine damage means more replacement parts.

For a given turbine inlet and condenser pressure, exhaust steam quality is controlled by the degree of superheat and turbine efficiency. We can still have high turbine efficiency and reasonable steam quality by upping the superheat. As the end-user, we have control over the degree of superheat and condenser pressure.

While it is easy to directly monitor EGT on a GT and thereby limit hot-section thermal damage, in the case of steam turbines we cannot directly measure the damage-causing agent...ie. LP-stage condensation. The only way to monitor exhaust steam quality is by performing a real-time energy balance across the turbine, which these days is not so difficult to do. We have the technology.

Somewhere between 80% and 100% exhaust steam quality there probably lies an economic optimum in terms of reliability, efficiency and life-cycle cost. But we can't perform this analysis until we can define even a crude mathematical relationship between LP-stage wear rate and exhaust steam quality. Perhaps, we could derive such a relationship by knowing the rate at which turbine efficiency falls off over time for a given set of steam conditions, and then assigning a reasonable portion of that degradation to observed post-run LP-stage damage. I don't have access to such data, but perhaps the public utility folks do. Perhaps the folks at EPRI or the US Navy might have a handle on this.

Perhaps the root-cause of most LP-stage damage is maloperation...inadequate superheat and pulling too hard on the condenser. Easy to do on a compressor drive where it is still uncommon to monitor turbine output torque/power directly. Again, we also have control over this by installing torque-measuring couplings and more-sophisticated turbine control systems.

Thank you for sharing your thoughts...and sweat...please don't overdo it on the treadmill!

Best regards,

Tom

 

[rmw]

tgmcg,

Can you give some specifics. 5MW isn't too big, but 14K rpm puts those machines in a class with the big boys with respect to their efficiency range.

Do you have a number for average efficiency of the typical machine?

I think you are on to the answer if you have the capability of raising the superheat of the inlet steam.

Most refinery mechanical drive turbines that I encounter long ago were pushed beyond their maximum, and the complaint is usually condenser back pressure problems, normally blamed on the air removal equipment or the condenser itself. (They never want to admit that they are trying to put a size 10 foot into a size 6 shoe.)

I'd like to do what I recommended to you and slide some straight edges up and down a mollier diagram, so if you can give me what steam conditions and CW temps you have to work with-what you have now, and what you might be able to go to-I might be able to give a more definitive answer.

rmw

 

[tgmcg]

rmw,

Most of the turbines are running off the 150# steam header at 150 psig & 320-360F. Condensers are operating at -26.5 inHg.

The question remains, how much water condensation should allow to keep LP-stage blade damage within an acceptable range. In order to justify any increase in steam header temperature, we have to be able to quantify the reduction of maintenance cost and sustained higher efficiency. It all boils down to economics. One cannot evaluate the economics without having both the costs and resulting savings associated with any proposed investment.

Best regards,

Tom

 

[rmw]

tgmcg,

I would guess that you are probably away from your office and giving those figures off the top of your head. Saturation temp for 150 psig is 365°F.

It would help to know what your upper limits are on raising the superheat on this 150# header.

But, that still gives me something to work with based on the pressure.

Regarding your economic justification, I have no idea about how to quantify the reduction of moisture, but recognize that any moisture at 14K rpm is lethal.

After I think about this, I will post back. (Heading to the treadmill-a good thinking place for me.)

Meanwhile, have you looked on;

http://www.gepower.com/prod_serv/products/tech_docs/en/all_gers.htm

and specifically under the link on steam turbines?

I think some of the GER's deal specifically with moisture related problems, but it has been a long time since I looked at them and I don't remember any of them doing it on a quantifiable basis.

More later (5 miles later).

rmw

 

[tgmcg]

rnw,

Yes, I'm away from my office, but have the turbine manufacturers data sheet here in my hand.

Normal steam inlet temperature = 150 psig
Normal steam inlet temperature = 366 F!
Exhaust pressure = 4.0 inHg Abs.

This is probably based on steam conditions supplied by the customer. The data sheets make no mention of exhaust stea, quality, which I calculate to be 89.6% given a turbine isentropic efficiency of 64% per current rated power output. I calculate Tsat = 366 F. Well that's just great, eh?

Note that the exisiting turbine MCOS is 10,486 rpm, and we are considering uprating it to 13,000-14,000 rpm...possibly a complete rotor, if not complete turbine, replacement. We need the speed for the re-rated compressor conditions. It is entirely possible we can't there from here and will convert to VFD motor drive.

Interesting how no mention is made in the datasheets about exhaust steam quality, or the total lack of inlet superheat. This is a perfect example of why you cannot depend on the vendor to keep you out of trouble.

Best regards,

Tom

 

[rmw]

Ahh! but you can get help here.

Katmar, one of the contributors to the site has a free software that I use for situations like this. It can be downloaded from

http://www.katmarsoftware.com

and is called turbine.

Using his software shows an exhaust quality of 89.62. Pretty close, huh?

Note the steam flow at this inlet condition.

Now, re-running that same turbine for 200°F superheat results in an exhaust quality of .980. Pretty high quality or low moisture.

Notice the steam flow at this revised inlet condition. 80K vs 91.5K using the 5MW you referenced above.

Now there is a quantifiable value for your economic evaluation if, and only if, the refinery is having to burn any fuel gas to generate process steam.

The fuel savings may or may not justify the project, and then the reduction of exhaust moisture is just lagniappe.

So, your moisture is virtually gone, you are saving fuel costs; the question is can you come up with 200 degrees of superheat in the 150# header.

Before you came back with your turbine information I had used a tool on the Dresser-Rand site that will size a turbine. Using the flow given in that result, you can iterate the Katmar Turbine program to find what efficiency that Dresser-Rand came up with.

Their program would not let me use 14K rpm. It limited me to 12K, which is close enough for this exercise. I have lost it now, but it was about what you came back with.

Re-running the D-R program with the SH produces a steam flow that iterates in the Turbine program to 72% efficiency and an outlet moisture of 95.41%.

So, the SH has increased the efficiency of the machine but the slope of the expansion line is not significantly different (64 vs. 72%). What is different, looking back to the mollier diagram, is that the beginning point is sliding up the pressure line to the 200 °F SH point, bringing the ELEP into a moisture region that meets your criteria.

I wouldn't go any higher in SH than 200 degrees because in hot weather if you can't maintain the 4"hga back pressure, then you would begin to exhaust superheated steam. Condensers don't like SH in the steam. Search this site for threads on that topic.

rmw

 

[tgmcg]

rnw,

Thank you for the helpful suggestions.

According to the manufacturers data sheet, the maximum allowable TIT is 400F, so the absolute best we'd be able to do is increase SH by 34F.

While this improves matters, it only gets us up to 91.3% exhaust steam quality....still not terribly encouraging in the absence of quantitative damage prediction guidelines.

Sure wish we had a more quantitative idea of what kind of damage and performance loss we might incur at 5%, 10% and 15% exhaust steam quality, respectively, over the course of a 3-year run.

With a compressor speed of 14,000 rpm and these steam conditions, I'm leaning toward converting all these compressors to electric motor drive. At least we will have a predictable result...that's what engineering is all about, IMHO. In this case, steam turbines probably aren't the right choice in terms of reliability and sustained efficiency. The turbines were first installed circa 1964 when efficiency and reliability weren't the high priority they are today. While they were probably the best available technology in terms of variable speed drives, we have better variable-speed drive technologies available to us today, without the unpredicatble long-haul performance issue.

Best regards,

Tom

 

[tgmcg]

rmw,

"... what kind of damage and performance loss we might incur at 5%, 10% and 15% exhaust steam quality..."

...make that 95%, 90% and 85% exhaust steam quality.

Regards,

Tom

 

[rmw]

tcmcg,

The performance loss can be calculated from the formulas I put in my first post. Delta ELEP is an enthalpy value, and you can convert that to performance or fuel costs.

Regarding maintenance costs, these things have been there since 1964 spinning at 10.5K rpm at a given percentage moisture at the outlet, so you should have a baseline for what the maintenance expense over the years for these rows is, if you can isolate the bottom end of the turbine.

I have seen factors posted in some of the threads on this site and other places for wear rates caused by moisture as a function of velocity. You might search the site. Your wear is a function of the blade speed, tip speed being the limiting factor and the moisture that they encounter as they spin.

I don't know of a given way to determine this wear. Others might, but they don't seem to be jumping in on this thread.

rmw

 

[norzul]

Dear tcmcg,

Yu mentioned the steam inlet conditions to the condensing ST are 150 psig(10.3barg)+ 366F(185.5C). This is almost saturated steam. Is this letdown steam or directly generated from the boiler?

The total electrical power required is 5MW X 10 = 50MW? What is the electricity tariff in your country?

 

[norzul]

Hi rmv,

I've downloaded the turbine-steam consumption calculator and still learning. I've posted the result of the calculation in my tech blog for the followings:
1) Inlet steam conditions : (150psig + 366F)
2) Exhaust steam conditions : (4 inHG abs)
3) Power output : 5 MW
4) Turbine efficiency : 64%

refer to :

http://technical2005.blogspot.com/

The outputs are:
1) Exhaust steam quality : 89.67%
2) Steam flowrate : 41.7 ton/hr

My questions:
1) The turbine efficiency as mentioned in the software, is it the turbine thermal efficiency or the turbine isentropic efficiency? What is the different?
2) How do we know it is 64%? Is it stated in the equipment datasheet?
3) Turbine thermal efficiency for this size (5MW) is probably less than 20%...is it correct?
4) With regards to exhaust steam quality, how low can the turbine accept?

Thanks

norzul

 

[tcs764]

Best efficiency can ussually only be achieved at the expense of raising the potential for last stage blade erosion. In my experience with larger units it has always been necessary to design for this by choice of blade materials, hard facings or the fitment of suitable erosion shields. It is also common practice to fit water extraction facilities before the last stage. Erosion damage is not in itself linear taking some long period to commence (ie due to initial hardening) then rapid as pitting occurs and then slows up ( ie droplets retained on eroded surface act as a buffer). Monitoring of the degree of erosion damage at successive major outages can provide a planned basis for maximising blade life and taking appropriate action should and when this is required.

 

[tgmcg]

tcs764,

Interesting comments. While sweeping generalizations can be troublesome, from a turbine design standpoint, do you have any thoughts as to the minimum recommended steam quality for sustained efficiency over a 3-5 year run? 95%? 90%? 85%? 80%? Or do you even think such a limit can or should be expressed? Feel free to provide any provisos you think necessary.

On smaller turbines (< 10 MW), it may not be so common to provide water drainage features.

Regards,

Tom

 

[tcs764]

Hi Tom
In my experience (on large units) the turbine manufacturer will have designed the unit to meet the specified terminal steam conditions, heat rate along with a specified range of cooling water temperatures/flows or exhaust pressures. As such the steam quality (in terms of wetness) is implicit in the design and should pose no problem as long as operation within the specifications is complied with. It is therefore not a parameter which in itself is used in the manner you suggest, it is also difficult to accurately evaluate or monitor on an operational unit.
Whilst steam quality (in terms of wetness) is a major factor in the erosion process other factors such as nozzle design, axial gap and the susceptability of the moving blading to erosion damage can determine whether any significant erosion occurs or not. Such factors will vary for each machine design and hence a generalised limitation on steam quality in this context may not be helpful.
Regards
tcs764