The Mechanisms of Performance Degradation in Gas Turbines

Christopher Haynes and James Koch www.heatate.com

Introduction The economics effect of degradation in a gas turbine power can be severe. Large combined cycle power can often experience a 4% loss of output and 2% worsening in heat rate. Such losses in a typical power market can be worth 5 million dollars per year. Diligent efforts to minimize these losses are warranted.

This paper summarizes a rational approach to minimizing degradation. It involves three elements. The first is quantifying and trending the degradation of the components that make up the gas turbine. The second is determining how much this degradation is impairing heat rate and capacity. The third, and perhaps most important, is to identify the physical changes in the gas turbine hardware that are causing the degradation.

Quantifying Degradation

It is well known that gas turbine performance degrades in service, with some of it being recoverable by cleaning and some recoverable during overhauls. The frequency and scope of these activities can be better planned if the degradation is trended using operating data.

Summary Parameters

Most commonly, a gas turbine’s’ performance is described by its four “Summary Parameters”. They are:

• Full load heat rate, • Full load output, • Full load air flow • Full load exhaust temperature.

The actual operating values of the four Summary Parameters can be determined using operating data. However, to determine if degradation has occurred they must be corrected to a standard external condition. When this is done, changes in performance caused by changes in external conditions will not be confused with changes due to degradation. Typically, the standard external condition might be 14.7 psia, 59 def F, 60% relative humidity, and design inlet and outlet pressure drops.

A disadvantage of trending the Summary Parameters is that they don’t indicate which element of the gas turbine is experiencing the degradation. Consider a case where the corrected output is trending downward. The corrected output trend doesn’t say why this is happening, or where the problem is. It might be due to problems in the compressor, the turbine section or a change in compression ratio.

Component Parameters

A different approach involves trending parameters that identify the component where the problem is. For that reason they are called Component Parameters. Most commonly, they are:

• Compressor Inlet Volume Flow • Compressor Efficiency • Turbine Inlet Flow Area • Turbine Expansion Efficiency

The Component Parameters describe the performance as completely as the Summary Parameters. The same operating data that is used to trend the Summary Parameters , (with a few modest additions such as the compressor discharge pressure and temperature), can be used to determine their actual operating values.

It would, of course be good if more detailed parameters, such as the efficiency of individual stages, or groups of stages could be obtained from plant data. Unfortunately this is not possible in a normally instrumented unit.

The Component Parameters can change when external conditions change, but the changes are small. The greatest correction is usually that for inlet temperature. However it can be seen in the table below, that the correction is only 1 or 2%, for compressor efficiency and volume flow respectively, while turbine inlet area and expansion efficiency do not change at all as ambient temperature changes.

As a result, corrections to standard conditions are either unneeded or, at most, rather undemanding in accuracy.

Typical Values of Component Parameters

at Different Inlet Air Temperatures

Typical Value

at 60 F Inlet Temp

Typical Value

at 20F Inlet Temp

Compressor Efficiency, % 88% 87%

Compressor Inlet Volume Flow, ACFM

1,000,000 980,000

Turbine Inlet Flow Area, sq inches 400 400

Turbine Expansion Efficiency, % 87% 87%

It would, of course be good if more detailed information about degradation, such as the degradation of individual stages or groups of stages, could be obtained from plant data. Unfortunately this is not possible in a normally instrumented unit.

The Effect of Degradation of Components on Plant Performance

The effect of component degradation on a power plant’s performance can be accurately determined by heat balance modeling. This has been done for a typical combined cycle power plant. One of the writers of this paper was the principle author of the CCYCLE heat balance program so this tool was used for this analysis. The power plant is shown on Figure 1, a CCYCLE diagram.

The heat balance modeling determined the effect of changes in Component Parameters on both the simple cycle and the combined cycle heat rate and output. These results are summarized below.

Effect of Changes in Component Parameters On Heat Rate and Megawatt Output

Component Parameter Change in Component

Parameter

Effect on Simple Cycle Output

Effect on Combined Cycle Output

Effect on Simple Cycle Heat Rate

Effect on Combined Cycle Heat Rate

Inlet Volume Flow 1% Decrease -1.0% -1.0% slight Slight

Compressor Effy 1% Decrease -1.8% -1.3% +1.6% +1.2%

Turbine Inlet Flow Area

1% Decrease -0.2% -0.3% -0.2% +0.2%

Turbine Expansion Effy

1% Decrease -2.6% -3.0% +2.2% +1.6%

One of the remarkable features of gas turbine power plants is how sensitive their performance is to changes in compressor and turbine efficiencies. This can be seen in the table above. A 1% decrease in both the compressor and the turbine efficiencies will result in a 4.3 % worsening in combined cycle heat rate. Thus in gas turbine plants, even modest amounts of degradation have serious consequences. By contrast, in conventional Rankine cycle steam units, a 1% decrease in all the sections of the steam turbine will worsen the unit’s heat rate by only 0.8%.

Turbine Expansion Efficiency

It will be noted that degraded turbine expansion efficiency has a very large effect on unit performance. There are two reasons. The first reason first is that the turbine section produces power for both the compressor and the generator. In a 200 megawatt gas turbine, the turbine section produces about 500 megawatts. A small change in efficiency of such a large power producer will have a big effect.

The second reason is that the method for controlling gas turbine firing temperature causes any degradation in turbine expansion efficiency to reduce firing temperature. Reducing firing

temperature will further worsen heat rate. And it reduces the heat input to the unit, causing a substantial reduction in output.

The gas turbine’s firing temperature control does not measure the firing temperature directly. Instead it estimates the firing temperature based on 1) the measured exhaust temperature, 2) the measured turbine pressure ratio, and 3) the design turbine expansion efficiency. If the true efficiency is different from the design value, perhaps due to degradation, the firing temperature will be controlled to the wrong value.

This can be seen by considering an extreme case, one where the true turbine expansion efficiency is zero. In such a case, the temperature drop through the turbine would be zero, at any pressure ratio. However, the control system would falsely believe that the normal temperature drop of about 1200 deg F, based on the design efficiency, was present. The control system would then under-fire the engine by 1200 deg F.

Examples of Typical degradation

(Later)

Mechanisms of Degradation

Although it is an obvious point, degradation is not all about test data and thermodynamic calculations. Degradation, in the final event, is about a physical change somewhere in the gas turbine. In other words, when degradation occurs, it is because something undesirable has happened to the gas turbine hardware. For this reason an essential aspect of minimizing degradation is a process of inspection of the gas turbine during overhaul and a reconciliation of the results of the inspection with the degradation. In this section we will examine the categories of degradation, their physical nature, how they can be determined and how they affect performance.

• Changes in Surface Roughness • Changes in Airfoil Shape • Changes leakage paths

Changes in Blade Surface Roughness

The effect of blade roughness can be found by consideration of basic theories for turbulent flow through pipe, which is then extended to plates and air foils.

Flow through a Pipe

The pressure drop needed to force flow though rough pipes can be found from a well understood theory familiar to engineers. The theory of flow over rough turbine and compressor blades shares many of the aspects of the familiar theory of pipes. For this reason, this section will discuss certain features of flow though a pipe, so they can be extended later to turbine and compressor blades.

Pipe Roughness, and Types of Turbulent Pipe Flow

The figures below show three cases of how surface roughness of the pipe affects turbulent flow.

In Case A, a laminar boundary layer is adjacent to the wall, and as shown, the wall roughness is smaller than this layer. The laminar boundary layer shields the flow in the turbulent region from feeling the roughness, so the flow through the pipe is as if the pipe were completely smooth. If the roughness were increased slightly, it would still be smaller than the laminar boundary layer, so flow would be unaffected.

In Case B, the wall roughness is greater, so that it begins to protrude outside the laminar layer and disturb the flow in the turbulent region. If the roughness were to increase any more, the flow would be disrupted and friction would increase. .

In case C, the wall is quite rough, and extends well outside the boundary layer. This is called “fully turbulent” flow. The rough protrusions disrupt and impede the flow in the turbulent region. Changes in the roughness would change the flow. If the roughness were to increase, the disruption and resistance to flow would be still greater. If the roughness were to decrease they would be less, until the roughness was within the boundary layer, when the flow would be the same as that for a completely smooth pipe.

Types of Flow Turbulent Flow in Pipe

Roughness and Pipe Losses

This resistance to flow in a pipe is quantified using the familiar Darcy equation:

The “friction factor”, f, is a dimensionless coefficient that defines the friction loss, when the velocity, density, and diameter are known. It was determined experimentally, and is given on Moody’s Chart, the figure below:

Moody’s Chart

The chart depicts how friction factor varies with Pipe Reynolds Number for various flow regimes. In the upper right, the friction factor is shown to slope down in a straight line in the region of laminar flow. When Reynolds number exceeds about 2000, there is a transition to turbulent flow. Notice that if the flow had remained laminar, and the line extrapolated, the friction factor would be much lower than in turbulent flow.

In the turbulent region, it can be seen that friction factor is a function two parameters. The first is the Pipe Reynolds Number. The second is the “Relative Roughness”, ( , which is the ratio of the height of the projections on the pipe wall to the pipe diameter. (The height of the bumps themselves, ε, is called the “Absolute Roughness”.)

In the upper right portion of Moody’s Chart, the lines of constant roughness are horizontal. This is the fully turbulent region, called “Case C” in the preceding discussion. Its boundary is the curving dashed line that is marked “completely turbulent” Notice that as the relative roughness lessens, say from 0.005 to 0.001, the friction factor lessens also (from 0.030 to 0.020)

To the left and below the fully turbulent region, the friction factor (for a given roughness) becomes greater as Reynolds Number decreases. In other words, as flow becomes more viscous, the resistance increases. For example, for a relative roughness of 0.005, the friction factor is 0.030 at Re = 200,000, while it is 0.038 at Re =10,000.

Threshold Roughness

Referring again to Moody’s Chart, notice the lines of constant roughness. They converge as Reynolds Number decreases, until they merge with the line of perfect smoothness. For example, consider a relative rougness of 0.005. When Reynolds number is less than 20,000, the friction factor is almost the same as that of a perfectly smooth pipe. This was “Case A” in the preceding discussion.

At any given Reynolds number, there is an amount of roughness below which friction factor is constant, and above which friction increases as roughness increases further. This is called the “threshold relative roughness”. When Re = 10,000 the threshold relative roughness is about 0.001, while when Re =10,000,000, the threshold relative roughness it is about 10-6.

Using data taken from the Moody Chart, a relationship between the threshold relative roughness and the Reynolds number for pipe flow is shown to be:

And the threshold absolute roughness in pipe is given by:

Notice that the threshold absolute roughness doesn’t depend on the pipe diameter. In other words, if the threshold absolute roughness is 1/1000th of an inch for 1 inch pipe, then it’s also 1/1000th of an inch for 24 inch pipe, carrying the same fluid at the same velocity.

The concept of threshold roughness would be important to a user of pipe, who wished to reduce pressure drop by polishing the inside pipe walls to make them smoother. He would gain no advantage by polishing the pipe below the threshold roughness. Further, this threshold roughness could be greater with viscous fluid, but it would be less for dense or fast moving fluids. Finally, as was just previously noted, the threshold roughness, and thus the polishing methods, would be the same for all pipe diameters.

It’s true that in the power industry pipe isn’t polished to reduce its pressure drop, but not because it wouldn’t work. It’s because it’s cheaper to reduce pressure drop by using bigger pipe. (Another reason, perhaps, is that rust or scale would quickly defeat the effect of polishing). However, the concept of threshold roughness will be seen to be very important in compressor and turbine blades. Their roughness can be reduced, at considerable cost, by polishing, cleaning etc. This has substantial value in improved performance, that is, until the surface finish has been improved to the threshold value. Further improvements have no value.

Flow over Compressor and Turbine Blades

Reynolds Number

It was seen that the features of flow through a pipe depended on Reynolds Number. Similar features of flow over blades also depend on Reynolds Number. As shown below, for pipe, there is one Reynolds number, based on diameter.

Reynolds Numbers For Pipes and Airfoils

For blades there are two Reynolds numbers. The first, called the “Blade Reynolds Number”, is based on the blade’s chord length. As an example of its use, when an airfoils lift and drag are tested, the results depend on the blade Reynolds number (as well as other parameters such as attack angle).

The second, called the Position Reynolds Number, is based on the distance from the leading edge. As an example of its use, when an airfoil is tested, it is found that the transition from laminar to turbulent flow depends on the position Reynolds number (as well as other parameters such as free stream turbulence and attack angle).

Flat Plate Boundary Layers

It has been found that the behavior of turbine and compressor blades can be usefully studied using the theory and data for boundary layers on flat plates. The figure below shows the development of a boundary layer as air flows over a flat plate.

As shown, the boundary layer begins very thin at the leading edge, widening as the flow continues along the plate.

At first the flow is laminar. After some distance, disturbances begin to disrupt the flow in a region of transition, until the boundary layer becomes turbulent. The boundary layer is

significantly thicker in the turbulent region. Immediately adjacent to the plate in the turbulent region there is a thin layer that is laminar.

Skin Friction on Turbine and Compressor Blades

The friction coefficient for a flow along a flat plate, developed from data of Prandtl and Sclichting, is shown on the figure below. The X axis is the position Reynolds number. What is striking is the similarity of the Prantl-Schlichting Chart to to Moody’s Chart. As in Moody’s Chart, there is a laminar flow region at low Reynolds Numbers. The friction coefficient for laminar flow is again much below that for turbulent flow.

Where the flow is turbulent, there are again two regions. The first is a region of complete turbulence, the second an intermediate region. In the region of completer turbulence the friction coefficient again does not vary with Reynolds number. In the intermediate region, the lines of constant relative roughness converge on the line of perfect smoothness as Reynolds number decreases.

Prandtl-Schlicting Chart

There are, however, some differences between the Moody and the Prandtl-Schlicting charts that may not be immediately apparent. In the latter, the Y axis is the local friction coefficient. It depends on position Reynolds number, the X axis. In other words, the friction coefficient changes along the length of the blade. Notice that for a “turbulent

smooth plate’, the friction coefficient decreases with Reynolds number, as it does for a turbulent smooth pipe. In any given section of pipe, the diameter is constant, so Reynolds number is constant, but for a blade, the Reynolds number increases along the chord length. Thus the friction coefficient is higher near the leading edge than it is farther back. The physics that explain this are shown on the figure below.

Boundary Layer and Velocity Profiles Along an Airfoil

Near the leading edge the boundary layer is thin and it widens along the chord length. Since the free stream velocity is more or less constant along the blade, the velocity profile will be relatively steep at the leading edge, with resulting high velocity gradients. Farther along the blade, the velocity profile is less steep, with more gentle velocity gradients. The amount of shear depends on the velocity gradient, and thus the amount of shear, and thus the friction, decreases along the chord length.

That friction decreases along the blade is also true for rough blades. As in the Moody Chart, the friction factor for a flat plate is seen to be constant in the fully turbulent region, regardless of Reynolds Number, provided the relative roughness is constant. However, this statement is somewhat misleading. In pipe, the relative roughness is defined relative to the pipe diameter, and for a given pipe the diameter doesn’t change along its length, so the relative roughness doesn’t change provided the absolute roughness is constant.

By contrast, for flat plates, the relative roughness is defined in terms of the distance from the leading edge. Thus, for a certain amount of absolute roughness, the relative roughness decreases along the length of the blade. As relative roughness decreases the friction coefficient decreases. This is shown on the figure below. The total friction drag along the first 10% of the blade is the same as on the last 30% of the blade.

Fluid friction Along the Length of an Airfoli

As can also be seen from the Prandtl-Sclichting Chart, the concept of concept of threshold roughness applies to flat plates, and thus by extension to turbine and compressor blades.

Using data taken from the Pranddtl-Schlicting chart, a relationship between the threshold relative roughness along a turbine blade and the Position Reynolds number is shown to be:

And thus the threshold absolute roughness for a blade is given by:

Notice that in the second formula, threshold absolute roughness doesn’t depend on the position along the chord of the blade. In other words, the threshold absolute roughness has the same value along the entire chord of the blade.

Defining and Measuring “Roughness”

Perhaps the most frustrating difficulty in discussing roughness is defining it. In the preceding discussion we described roughness, rather vaguely, as the “height of the projections” on the surface. This isn’t very useful, because the projections that comprise the roughness can vary in density, nature, and size distribution.

Surface finish is often defined and measured as the center line average (c.l.a.) of the departure from a flat surface, obtained by a stylus trace. This c.l.a. method has the advantage of being objectively measureable and rigorously defined. It has, however a grave disadvantage. It doesn’t correspond well to what we are interested in, the

disruption and resistance to flow caused by roughness. Consider the two surfaces shown below. They have projections of the about same height, and will give about the same c.l.a, but the one with the sharp projections has substantially more resistance to flow than the one with rounded ones.

Different Types of Surface Roughness with the same Height of the Projections

An effective and practical approach to defining roughness, for the purpose of studying flow resistance, is to define roughness in terms of sandpaper fineness grades. (Sandpaper is finer as the grade number increases.) This allows flow data on the effect of roughness to be consistently obtained in laboratories, where sandpaper of various grades can be glued to surfaces. It also allows for easy, if subjective, measurement of the roughness of surfaces encountered in shops and in the field, by comparing the “feel” of the surface with the feel of samples of sandpaper of various grades. The standard sandpaper grades are cheap and readily available at most car parts stores. Cotton, Kindl, Chioffi and others at Encotech, Inc developed and successfully used this approach for steam turbine efficiency audits. The present authors are applying it to gas turbines.

Most of the classical early work on the effect of roughness on flow, done by Nikuradse and others, used a related but different method. It involved gluing tightly packed grains of sand, of fixed size, onto surfaces. (The grains of sand on sandpaper are not of uniform size, Instead their sizes conform to a specified grain size statistical distribution) There is a large amount of flow data available this basis. In actual flow tests on flat plates, Speidel compared the results of this fixed sand grain size method, with results of the sandpaper grade method. This comparison allows the flow data from various sources to be used interchangeably. The table below is based on his results. It gives the sandpaper grade and the uniform sand grain size that yield the same flow resistance.

Sandpaper Grade and Sand Grain Size that yield Identical Flow Resistance

Sand Paper Grade

(CAMI)

Uniform Sand Grain Size , mils

80 11

100 7.1

220 4.0

320 2.4

600 0.90*

1200 0.36*

1500 0.30*

2000 0.20*

3000 0.13*

*extrapolated from Speidel’s data (one mil is 1/1000th of an inch)

The table above uses the Coated Abrasive Manufacturers Institute (CAMI) designations. (CAMI has merged with other abrasives industry organizations, and is now the Unified Abrasives Manufacturers' Association) The CAMI standards are prevalent in the US. In most other countries, the Federation of European Producers of Abrasives (FEPA) standards are prevalent.

The potential for confusion between CAMI and FEPA is substantial. The grade numbering systems look similar. CAMI grades are strictly numbers, such as 220, 320, 1200, while FEPA grades begin with the letter P, such as P220, P320, P1200, etc. In spite of the similar looking designations, the corresponding grain sizes differ substantially, for grades above 320. For the same grade number, the FEPA grade are coarser. For example, a CAMI grade 1200 has a similar grit size and roughness of a FEPA grade P2500.

The authors of the present paper have been informed that some imported papers sold in the US use the FEPA standard, but without the letter P. A person who purchased such paper marked 2500, would assume it was CAMI 2500, and thus believe that the uniform sand grain size was about 16 mils, when it was actually 36 mils. The 3M Company is said to always use the proper designation.

The Effect of Roughness on Velocity Profile in the Boundary Layer

Later

Effect of Roughness in a Typical 200 MW Power Generation Gas Turbine

The effect of roughness has been evaluated for a 200 MW Class Gas Turbine Engine. The table below gives the relevant physical parameters for this machine at four locations:, the first and last stages of the compressor, and the first and last stages of the turbine. It can be seen that the blade Reynolds number is more or less constant through the compressor, while the blade numbers are lower in the turbine, and especially low at the exhaust.

Selected Parameters for a 200 MW Gas Turbine Engine

Threshold Roughness in the 200 MW Engine

The table below gives the threshold roughness for the same locations in a the 200 MW turbine. It can be seen that the threshold roughness for the first stage compressor blades is CAMI sandpaper grade 1500. Highly polished blades can attain better smoothness, but it has no value.

For the last stage compressor blades, the corresponding sandpaper grade is CAMI 4500, which is beyond practical possibility. Thus any improvement to the smoothness of these blades will be beneficial.

Finally, the threshold roughness of the last stage of the gas turbine is CAMI 500. This is a rather coarse finish. There is no benefit to making these blades smoother.

Threshold Absolute Roughness for a 200 MW Engine

Application of Theory to the Compressor

As we have seen, skin friction increases when blade roughness increases above the threshold value, the being shown on the Prandtl Schlichting Chart. In a compressor, the increased skin friction results diminished efficiency and flow.

Effect of Roughness on Compressor Efficiency

The authors have estimated these losses for compressor and determined the resulting amount of diminished efficiency. It was assumed that both rotating and stationary blades had the same amount of roughness. Some results of this estimate are shown on the figure below, for the first and last stages. The base relative efficiency, 100%, is that attained with perfectly smooth blades

Compressor Stage Efficiency vs. Blade Roughness

for a 200 MW Engine

As previously noted, the threshold roughness is corresponds to 1500 and 4500 grade paper for the first and last stages respectively. These values can be seen on the figure above, as the points where the graphs change slope. Once the threshold roughness is reached, further changes in roughness affect both stages about equally. A worsening of the roughness of all stages from 1500 grade to 800 would reduce compressor efficiency by about 1%, and thus worsen heat rate and megawatt output by about 1.5%.

Effect of Roughness on Compressor Air Flow

Later

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