DESIGN PROJECT: DESIGN OF A GAS TURBINE

ME I3100 STEAM & GAS TURBINES

Fall 2017

DESIGN PROJECT: DESIGN OF A GAS TURBINE

Submitted to: Prof. Rishi Raj

Submitted by: Carles Bertran

EMPLID: 23506295

December 5th, 2017

M.E. I3100 Steam & Gas Turbines DESIGN PROJECT

ABSTRACT

This paper collects the data and calculations necessary for the design of a gas turbine.

A gas turbine is an internal combustion engine that uses air as the working fluid, and is

composed by an upstream rotating compressor coupled to a downstream turbine by a

combustion chamber. The application of the turbine of this paper is to generate

electricity for a power plant, by transforming the thermal energy of a fuel into

mechanical energy. This mechanical energy will be used to power a generator which

transforms mechanical energy into electrical energy. The final design has been

determined by the design parameters that have been preset according to the

requirements of the application of the turbine.

In order to transform the energy of fuel into mechanical energy (Work), the gas turbine

engine follows the thermodynamic cycle known as the Brayton cycle. This cycle has

been analysed in depth for the ideal and real Brayton gas turbine cycles, to calculate

the parameters that govern the design of the turbine. These parameters have helped

design the turbine by determining the dimensions of the components of the

turbomachine. Calculations determined that this turbine must have four stages with high

stage enthalpies. Once the parameters have been met, the following chapters describe

the design and definition of the component materials chosen for the design, as well as

the bearings required for the assembly.

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TABLE OF CONTENTS

1. BACKGROUND 3

2. THEORY OF OPERATION THE BRAYTON CYCLE 5 2.1. IDEAL BRAYTON CYCLE PROCESSES 6

3. CYCLE ANALYSIS AND CALCULATIONS 7 3.1. SYMBOLOGY AND PARAMETERS 8 3.2. IDEAL CYCLE 9

3.2.1. SUMMARY OF STATE PARAMETER CALCULATIONS 10 3.2.2. WORK and ENERGY 10

3.3. REAL CYCLE 11 3.3.1. SUMMARY OF STATE PARAMETER CALCULATIONS 13 3.3.2. WORK and ENERGY 13

3.4. TURBINE CALCULATIONS 14 3.4.1. STAGE CALCULATION 14 Calculation of STAGE 1 15 3.4.2. SUMMARY OF STAGE CALCULATIONS 19

4. SELECTION OF MATERIALS AND BEARINGS 20 4.1. MATERIAL SELECTION 20 4.2. BEARING SELECTION 23

4.2.1. ROLLING BEARINGS 24 4.2.2. JOURNAL BEARING 25

5. DISCUSSION 28

6. REFERENCES 29

2


1. BACKGROUND

The gas turbine, also known as combustion turbine, is a type of internal combustion

(I.C.) engine. To differentiate it from the I.C. piston engine, each one of the

thermodynamic processes of the turbine occur simultaneously throughout different

elements of the equipment. For instance, it uses an upstream axialflow rotating

compressor that increases the pressure of air at the same time that this one is being

returned to atmospheric pressure after providing work on the downstream turbine. In

I.C. piston engine, an oscillating piston has the role of compressing and providing work

on sequential stages.

In gas turbines, the compressed air is driven into the combustion chamber where fuel is

mixed in and ignited. The increase in pressure after ignition is used to move the

downstream turbine at the end of the engine which, in most cases, is used to drive the

compressor as well as other mechanical elements like, in power generation, an electric

generator coupled to the shaft of the turbine. In airbreathing jet engines however, the

high pressure of fluid after ignition is known as the fastmoving gases or jet discharged

at the exhaust which generate the propulsion (thrust) necessary to propel aircrafts at

supersonic speeds. On occasions, the high temperature of the ignition is also used at

the exhaust (waste heat) to feed a regular steam generator that powers a steam turbine

(combined cycle), which can reach efficiencies of over 60%. However, this paper

focuses on the main piece of equipment of the gas turbine like the one shown in Figure

1.

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Figure 1: Basic components of the gast turbine

http://cset.mnsu.edu/engagethermo/images/gasturbineanimation.png

The application of turbines is very versatile: in power generation they can be used to

operate additional emergency generators to supply energy during peak demand, due to

the ability to be quickly turned on and off compared to other forms of power generation.

Its relatively simple mechanism allows the construction of a wide range of sizes and

weights, which further extends its applications. This mechanical simplicity also makes

turbine engines especially suited for high power outputs during an extended period of

time, and are known to be very reliable engines with very long operational life under

extended hours of use. These features explain its wide use in the modern aeronautic

industry and other sorts of transportation.

Simple cycle gas turbines are still not as efficient as other power generating engines,

therefore their use are still limited to emergency situations or mostly military applications

where to benefit from its high power generation.

4

http://cset.mnsu.edu/engagethermo/images/gasturbineanimation.png


2. THEORY OF OPERATION THE BRAYTON CYCLE

Figure 2. Schematic of the gas turbine. The numbers show where the cycle processes occur.

Gas turbines follow a thermodynamic cycle called Brayton cycle. In this cycle there is a

sequence of thermodynamic processes which describe the transfer of heat and work

into and out of the system while varying pressure, temperature, volume and enthalpy,

and which return the system to its initial state closing the thermodynamic cycle. In this

paper the ideal and actual Brayton cycles were analysed, therefore we discuss in detail

the processes that compose these two cycles:

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2.1. IDEAL BRAYTON CYCLE PROCESSES

Figure 3. The Brayton Cycle: Pressure Volume diagram (Left) and Temperature Entropy

diagram (Right).

12. Isentropic Compression

Air at atmospheric pressure and temperature enters the compressor where it gets

compressed. Work is added to the process (Wc). This is an adiabatic, reversible (thus

isentropic) process.

23 Isobaric Process

Compressed air runs through the combustion chamber where is mixed with fuel and

ignited. This process adds heat to the cycle but the pressure remains constant because

the combustion chamber is open to flow in and out.

34 Isentropic Expansion

High heat, high pressure air returns to atmospheric pressure (adiabatic reversible

expansion) expanding and producing work, which is mostly absorbed by the turbine

(Wt).

41 Isobaric Process

Air exits the turbine (exhaust) rejecting heat to the atmosphere where it returns to initial

values of temperature and pressure.

6


3. CYCLE ANALYSIS AND CALCULATIONS

In this section of the paper the parameters for the thermal cycle have been calculated

from the design parameters that have been set. The requirements from the application

define the initial parameter such as the shaft speed and power needed from the turbine.

The rest of the initial parameters are set by physical constraints like the maximum

temperature that the available materials are capable of withstanding, or the

environmental conditions in which the engine will be operating on:

Parameter Symbol Value

Temperature at Turbine Inlet (Max. Temp.) T 3 1800 °F = 2260 °R

Temperature at Compressor Inlet (atmospheric) T 1 70 °F = 530 °R

Pressure at Comp. Inlet (atmospheric) P 1 1 atm = 14.69 psi

Net Shaft Power P NET 10000 HP = 7457 KW

Shaft Speed N 7200 RPM

Table 1. Initial/design Parameters

By analysing the Brayton thermodynamic cycle, the values of pressure, volume,

temperature, enthalpy, mass flow rate and work can be defined at every process. This

analysis has been done for the ideal cycle as well as the real cycle.

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3.1. SYMBOLOGY AND PARAMETERS

p Pressure (PSI)

T Temperature (°R)

h Enthalpy (Btu/lbm)

q in Heat addition (Btu/lbm)

q out Heat rejection (Btu/lbm)

W Work (Btu/lbm)

ṁ Mass Flow Rate (lbm/s)

W c Work of compressor (Btu/lbm)

W T Work of Turbine (Btu/lbm)

P Power (KW)

N Number of cycles (rpm)

N b Number of blades

V Absolute velocity (ft/s)

W Relative velocity (ft/s)

U Axial velocity (ft/s)

C p Specific heat at constant p (Btu/lbm °R)

g c Massforce constant (ld ft/lbfs 2 )

k Ideal Gas Constant (Btu/lbm°R)

r p Pressure ratio

η Efficiency

l length of blade(ft)

r m Middle radius (ft)

r t Tip radius (ft)

r h Hub radius (ft)

ρ Density (lbm/ft 3 )

φ Loading coefficient

ϕ Zweifel Factor

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3.2. IDEAL CYCLE

Recall the following values for an ideal cycle:

K = 1.4

Cp = 0.24 Btu/lbm °R

air = 0.0748 lbm/ft 3ρ

State 1

Assume atmospheric T and P:

atm 4.69 psip1 = 1 = 1

0°F 30 °RT 1 = 7 = 5

T .24 40 27.20 Btu lbm h1 = Cp 1 = 0 ∙ 5 = 1 /

State 2

094.44°R T 2 = (T T )1 31 2/ = √530 260∙ 2 = 1

4.69 85.88psi p2 = p1 ∙ ( T 1

T 2 )k (k−1)/

= 1 ∙ ( 5301094.44)1.4 0.4/ = 1

T .24 094.44 62.66 Btu lbm h2 = Cp 2 = 0 ∙ 1 = 2 /

State 3

85.88 psip3 = p2 = 1

260°RT 3 = Tmax = 2

T .24 260 42.4 Btu lbm h3 = Cp 3 = 0 ∙ 2 = 5 /

State 4

4.69 psip4 = p1 = 1

260 094.44°R T 4 = T 3 ∙ ( p3p4 )

k−1 k/= 2 ∙ ( 14.69

185.88)0.4 1.4/ = 1

T .24 094.44 62.66 Btu lbm h4 = Cp 4 = 0 ∙ 1 = 2 /

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3.2.1. SUMMARY OF STATE PARAMETER CALCULATIONS

State Temperature ( )R ° Pressure (psi) Enthalpy (Btu/lbm)

1 530 14.69 127.20

2 1094.44 185.88 262.66

3 2260 185.88 542.4

4 1094.44 14.69 262.66

Table 2. Ideal Cycle state calculation results

3.2.2. WORK and ENERGY

Heat input

42.4 62.66 79.73 Btu lbm qin = h3 − h2 = 5 − 2 = 2 /

Heat rejected

62.66 27.20 34.46 Btu lbm qout = h4 − h1 = 2 − 1 = 1 /

Turbine Work

42.4 62.66 79.74 Btu lbm W T = h3 − h4 = 5 − 2 = 2 /

Compressor Work

62.66 27.20 35.46 Btu lbm WC = h2 − h1 = 2 − 1 = 1 /

Net Work

79.74 35.46 44.26 Btu lbm W net = W T −WC = 2 − 1 = 1 /

Thermal Efficiency

.5173 00 1.57% ηtheoretical = qinW net = 279.73

144.72 = 0 × 1 = 5

Pressure Ratio

2.65rp = p1p2 = 14.69

185.88 = 1

Mass Flow Rate (ṁ)

P net = 3412ṁ∙Δh

h h ) h ) Δ = ( 3 − h4 − ( 2 − h1

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0000 HP 457 kWP net = 1 = 7

75810.42 lbm hr 8.98 lbm s ⇒ṁ = P ∙3412net(h −h )−(h −h )3 4 2 1

= 7457∙3412279.74−135.02 = 1 / = 4 /


Heat Input q in 279.73 Btu/lbm

Heat Rejected q out 135.46 Btu/lbm

Compressor Work W c 135.46 Btu/lbm

Turbine Work W T 279.74 Btu/lbm

Net Work W net 144.27 Btu/lbm

Cycle Efficiency ηtheoretical 51.57%

Pressure Ratio r p 12.65

Mass Flow Rate ṁ 48.98 lbm/s

Table 3. Ideal cycle parameter calculation results.

The results from the calculations validate the design parameters in an ideal situation. A

high efficiency has been achieved. The following section determines the same

parameters for the real cycle.

3.3. REAL CYCLE

In this section the efficiencies of the components of the engine have been considered.

In other words, the elements of the assembly have energy losses due to mechanical

inefficiencies and other physical conditions. These have been assumed from examples

found in real life, and are presented at the beginning of the calculations.

In order to simplify the calculations, the values of K, C p and C v have been considered

constant, as in the ideal cycle. Thus the calculation process is presented as follows:

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Assumptions:

= 0.85ηcompressor

= 0.93ηturbine

State 1

Following the same fashion as for the

ideal cycle, assume atmospheric T and P:

atm 4.69 psip1 = 1 = 1

0°F 30 °RT 1 = 7 = 5

T .24 40 27.20 Btu lbm h1 = Cp 1 = 0 ∙ 5 = 1 /

State 2

30 194.05°RT ′2 = T 1 +(T −T )2 1

ηcompressor= 5 + 0.85

1094.44−530 = 1

4.69 52.14 psi p′2 = p1 ∙ ( T 1

T ′2 )k (k−1)/

= 1 ∙ ( 5301194.05)1.4 0.4/ = 2

27.2 86.05 Btu lbmh′2 = h1 +(h −h )2 1

ηcompressor= 1 + 0.85

262.22−127.2 = 2 /

State 3

52.14 psip′3 = p′2 = 2

260°RT 3 = Tmax = 2

T .24 260 42.4 Btu lbm h3 = Cp 3 = 0 ∙ 2 = 5 /

State 4

4.69 psip4 = p1 = 1

(T ) 260 .93(2260 094.44) 176.03 psi T ′4 = T 3 − ηturbine 3 − T 4 = 2 − 0 − 1 = 1

(h ) 42.4 .93(542.4 62.22) 81.83 Btu lbm h′4 = h3 − ηturbine 3 − h4 = 5 − 0 − 2 = 2 /

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3.3.1. SUMMARY OF STATE PARAMETER CALCULATIONS

State Temperature ( )R ° Pressure (psi) Enthalpy (Btu/lbm)

1 530 14.69 127.20

2 1194.05 252.14 286.05

3 2260 252.14 542.4

4 1176.03 14.69 281.83

Table 4. Real Cycle state calculation results

3.3.2. WORK and ENERGY

Heat input

42.4 86.05 55.82 Btu lbm q′in = h3 − h′2 = 5 − 2 = 2 /

Heat rejected

81.83 27.20 55.05 Btu lbm q′out = h′4 − h1 = 2 − 1 = 1 /

Turbine Work

42.4 81.83 60.15 Btu lbm W ′T = h3 − h′4 = 5 − 2 = 2 /

Compressor Work

86.05 27.20 59.37 Btu lbm W ′C = h′2 − h1 = 2 − 1 = 1 /

Net Work

60.57 58.85 00.78 Btu lbm W ′net = W ′T −W ′C = 2 − 1 = 1 /

Thermal Efficiency

.3968 00 9.68% ηreal = q′inW ′net = 256.35

101.72 = 0 × 1 = 3

Pressure Ratio

7.16r′p = p1p′2 = 14.69

252.14 = 1

Mass Flow Rate (ṁ)

P net = 3412ṁ∙Δh

h h ) h ) Δ = ( 3 − h′4 − ( ′2 − h1

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0000 HP 457 kWP net = 1 = 7

52460 lbm hr 0.12 lbm s ⇒ṁ′ = P ∙3412net(h −h )−(h −h )3 ′4 ′2 1

= 7457∙3412260.57−158.85 = 2 / = 7 /


Compressor efficiency ηcompressor 0.85

Turbine efficiency ηturbine 0.93

Heat Input q’ in 255.82 Btu/lbm

Heat Rejected q’ out 155.05 Btu/lbm

Compressor Work W’ c 159.37 Btu/lbm

Turbine Work W’ T 260.15 Btu/lbm

Net Work W’ net 100.78 Btu/lbm

Cycle Efficiency ηreal 39.39%

Pressure Ratio r’ p 17.16

Mass Flow Rate ṁ’ 70.13 lbm/s

Table 5. Real cycle parameter calculation results.

The efficiency for the real cycle is 39.39%, which is a significant drop in respect to the

ideal situation (51.73%). This is translated into a higher flow rate in order to achieve the

same max. Power. Since this scenario is a more disadvantageous case, these

parameters and results are being used to proceed with the design of the engine.

3.4. TURBINE CALCULATIONS

3.4.1. STAGE CALCULATION

At this point of the calculation process it is possible to determine the total enthalpy

change, which corresponds to the work done by the turbine. Since the typical values of

14


enthalpy change range between 50 80 BTU/lbm, the total enthalpy has to be divided in

different stages that fit the range. This defines the number of stages of the turbine:

Total Enthalpy Change

h ) 60.15 Btu lbm Δhtotal = ( 3 − h′4 = W ′T = 2 /

Enthalpy per stage 0 Btu lbm 50 h≤ Δ stage ≤ 8 /

5.04 Btu lbm Δhstage =Δhtotal

# of stages = 4260.57 = 6 /

This value of enthalpy is the average value through the 4 stages. For this design the

enthalpy per stage must follow a decreasing value from the first stage to the last. These

values have been adjusted conveniently for the calculation. Similarly, the values for the

reaction follow the same criteria but on an increasing form.

This parameters allow to start with the calculation of all the parameters that define the

design of the turbine in each one of its stages. The following part will walk the reader

through the calculation process of the parameters and dimensions that define the first

stage of the turbine. The results have been collected in table 6, where the values of the

results of the calculations for Stages 2 to 4 can be found as well.

Calculation of STAGE 1

1. (Specified)8.04 Btu lbm Δhstage1 = 6 /

2. Reaction at Stage 1 R = 0.32 (Specified)

R = Δhstage

Δhrotor

h h .32 8.04 1.77 Btu lbm Δ rotor = R ∙ Δ stage = 0 ∙ 6 = 2 /

h h h 8.04 1.80 6.26 Btu lbm Δ stator = Δ stage − Δ rotor = 6 − 2 = 4 /

15


3. Adiabatic Velocity

V o =√2 h 78∙ gc ∙ Δ stage ∙ 7 =

840.59 f t s = √2 2.2 8.04 78∙ 3 ∙ 6 ∙ 7 = 1 /

4. Velocity Ratio .60 UV o≈ 0

(Specified for a R = 0.32)

5. Blade velocity

.60 840.59 .60 104.35 f t s U = V o ∙ 0 = 1 ∙ 0 = 1 /

6. Mean radius ‘r m ’

U = 602∙π∙N ∙rm

.47 f t ⇒ rm = U ∙602∙π∙N = 2∙π∙7200

1104.35∙60 = 1

7. Absolute velocity

V 2 =√2 1 ) h 78∙ gc ∙ ( − R ∙ Δ stage ∙ 7 =

= √2 2 1 .32) 8.04 78∙ 3 ∙ ( − 0 ∙ 6 ∙ 7 =

517.79 f t s = 1 /

8. (Specified) α2 = 74o

9. rctan β2 = a ( V cosα2 2

V sinα −U2 2 ) = rctan 0.29 = a ( 1517.79 ∙ cos 74 o

1517.79∙sin 74 −1108.62o ) = 4 o

10. Relative Velocity

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48.44 f t s W 2 = cos(β )2V cos(α )2 2 = cos(40.29 )o

1517.79∙cos(74 )o = 5 /

11. Absolute Velocity

os(α ) 523.66 os (74 ) 18.36 f t s V 2X = V 2 ∙ c 2 = 1 ∙ c o = 4 /

12. Relative Velocity

W 3 =√2 h 78∙ gc ∙ Δ rotor ∙ 7 +W 22 =

176.81 f t s = √2 2 1.77 78∙ 3 ∙ 2 ∙ 7 + 548.442 = 1 /

13. rcos β3 = a ( W 3

W cosβ2 2 ) = rcos 9.17 = a ( 1176.81

548.44∙cos 40.29 o ) = 6 o

14. rctan α3 = a ( V cosα2 2

U−W sinβ3 3 ) = rctan .60 = a ( 1517.79 ∙ cos 74 o

1104.35−1176.81∙sin 69.17 o ) = 0 o

15. Absolute Velocity V 3

18.38 f t s V 3 = cos(α )3V cos(α )2 2 = cos(0.60 )o

1517.79∙cos(74 )o = 4 /

16. Work

W = Ug ∙778c

∙ (W inβ inβ )2 ∙ s 2 +W 3 ∙ s 3 =

4.52 Btu lbm = 32∙7781104.35 ∙ (548.44 in 40.29 176.81 in 69.17 )∙ s o + 1 ∙ s o = 6 /

17. Stage efficiency

.9483 00 4.83% ηstage1 =WV 2o

2∙778∙gc

= 64.52

2∙778∙321840.352

= 0 ∙ 1 = 9

18. Length of blades ‘l’ (rotor and stator)

r ) π ṁ′ = ρ ∙ V 2X ∙ A = ρ ∙ V 2X ∙ π ∙ ( t2 − r2h = ρ ∙ V 2X ∙ 2 ∙ rm ∙ l

17


⇒ l = ṁ′ρ∙V ∙2π∙r2X m

Recall , therefore0.13 lbm s ṁ′ = 7 /

.2435 f tl = 70.130.0748∙418.36∙2π∙1.46 = 0

19. Number of blades – Zweifel relation ( ).85 ϕ = 0

os β ϕ = b2∙S ∙ tan β an β[ 2 + t 3] ∙ c 2

3

We know that, π N b ∙ S = 2 ∙ rm

Given that the range of the

values for the cord length ‘c’

must be between 23 inches, for

this report it has been preset as

3 inches. Then,

.333 f t b ≈ c = 0

⇒ S = ϕ∙b2∙ tan β +tan β ∙cos β[ 2 3] 2

3=

.322 f t= 0.85∙0.3332∙ tan 40.29 +tan 69.17 ∙cos 69.17[ o o] 2 o = 0

Number of blades:

8.57 9 blades N b = 0.3222π∙rm = 0.322

2π∙1.46 = 2 ⇒ N b = 2

20. Loading Factor

.317 0.8 .317 φ = WU2

778∙gc

= 64.52

778∙321104.352

= 1 → < 1 < 2

This is an acceptable value for the R in this stage ⇒

The results for the calculations of the stages 24 have been collected on table 6.

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3.4.2. SUMMARY OF STAGE CALCULATIONS

STAGE 1 STAGE 2 STAGE 3 STAGE 4

Δh stage (Btu/lbm) 68.04 66.04 64.04 62.04

R 0.32 0.34 0.36 0.38

Δh rotor (Btu/lbm) 21.77 22.45 23.05 23.57

Δh stator (Btu/lbm) 46.27 43.59 40.98 38.46

V 0 (�/s) 1840.59 1813.33 1785.66 1757.56

U/V 0 0.60 0.61 0.63 0.65

U (�/s) 1104.35 1106.13 1124.97 1142.41

N (rpm) 7200 7200 7200 7200

r m (�) 1.47 1.47 1.49 1.52

V 2 (�/s) 1517.79 1473.16 1428.53 1383.90

α 2 (°) – assumed 74.00 76.00 78.00 80.00

β 2 (°) 40.29 42.21 42.52 42.53

W 2 (�/s) 548.45 481.16 402.97 326.12

V 2x (�/s) 418.36 356.39 297.01 240.31

W 3 (�/s) 1176.81 1161.68 1144.67 1131.45

β 3 (°) 69.18 72.13 74.96 77.74

α 3 (°) 0.61 0.08 3.76 8.70

V 3 (�/s) 418.38 356.39 297.65 243.11

W (Btu/lbm) 64.52 63.49 62.26 60.85

η stage (%) 94.83 96.14 97.22 98.10

m DOT m ̇ (lbm/s) 70.13 70.13 70.13 70.13

L (�) 0.24 0.29 0.34 0.41

b=3 inch (�) – assumed 0.33 0.33 0.33 0.33

c=3 inch (�) – assumed 0.33 0.33 0.33 0.33

Zweifel rel. Φ – assumed 0.85 0.85 0.85 0.85

S (�) 0.32 0.38 0.45 0.57

N b (blades) 29 25 21 17

Loading Factor φ 1.32 1.29 1.22 1.16

Table 6. Results for the calculations of the parameters that define the 4 stages of the turbine

19


4. SELECTION OF MATERIALS AND BEARINGS

4.1. MATERIAL SELECTION

In this section the material selection in gas turbine engine design is discussed. Material

selection in gas turbine is an important subject due to the large variety of requirements

needed to maximize the performance and operability of the engine. During the different

processes of the thermal cycle, the materials that compose the elements of the engine

are subject to very different and extreme situations, pushing the limits of the most

sophisticated materials available today. The challenges faced by these materials are

high loads, high vibrations (fatigue resistance), high impacts, contact with abrasive

particles like sand and other objects found in the air that enters the inlet, contact with

oils, oxidation and corrosion, high temperatures and or large temperature gradients.

Therefore, the materials selected will have to excel primarily in strength, fatigue

capability, weight, cost, creep resistance and thermal stability (low coefficient of thermal

expansion).

Let’s take a closer look at the specific elements throughout the engine, with the intention

to analyse the situations in which the materials have to perform. Once we have a basic

idea of these conditions, the requirements of the materials can be determined. As a

result we will be able to list some of the materials that are most suitable for each

situation.

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FAN

At the inlet, the blades of the fan are subject to high tensile strength due to the

centrifugal forces during the rotation of the shaft. Centrifugal forces can be minimized

with a lightweight material, which needs to be able to resist high impacts from outside

objects being pulled inside the engine.

Examples: Polymer Composite of Titanium alloys for blades, and nickelbased alloys

and polymer composites or titanium alloys for the containment of the fan.

COMPRESSOR

The compressor operates at high temperatures during long periods of time. Like the fan

and most of elements in the engine, the rotation of the blades occur at very high

speeds. Therefore the material is required to withstand high stresses, centrifugal forces

and fatigue.

21


Examples: Titanium alloys for blades operating at lower temperature, and nickel based

or titanium alloys on higher temperatures. Same for compressor disk.

COMBUSTOR

Very high temperatures (averaging at 2800F) during very long periods of time. The

combustor lining must be able to withstand stresses due to heat, takeoff and cool down

situations, and have good oxidation resistivity.

Example: Nickel based alloys for combustor and liner.

TURBINE

The materials required for this element must have good rotational strength, resist

pressure loading (from the very high pressure gases coming from the combustor),

withstand high temperatures, resist creep and oxidation.

Example: Single crystal Nickel based alloys with thermal barrier coatings.

MIXER

Materials of the mixer must withstand high temperatures as well as low temperatures,

requiring a material that has minimal effect towards thermal expansion.

Example: Nickel based alloys.

NOZZLE

At the nozzle the materials must be able to withstand high temperatures, with ranges of

1200F to 2400F.

Example: Nickel or titanium based alloys or ceramic matrix composites.

The following table lists some of the alloys currently employed in turbine engine

development. This is only a sample of materials used in some of the components found

in turbines which are capable of withstanding high temperatures without losing

mechanical properties. However, the list of materials from other families other than

22


alloys are extensively used in the design of modern turbines where the ratio of high

temperature resistance capabilities to strength are not as crucial, like ceramic matrix

composites, or polymer matrix composites for the use in combustors.

Other lightweight materials like carbon fiber reinforced composites are used for static

engine structures and inlet blades where significant weight savings result in more

efficient engines.

Table 7. Common Alloys used in the elements that compose turbine engines, with composition

percentages.

4.2. BEARING SELECTION

Gas turbines are a type of engine that shines for its mechanical simplicity. This simplicity

is mainly due to all the moving parts (compressor and turbine) having only a rotational

movement and being concentric to the same shaft. In order to guarantee a high

23


performance for the elements of the engine, a big responsibility is entrusted to the

bearings which allow rotational movement of the moving parts.

Different bearings are strategically positioned along the lengths of the shaft. Radial

support is provided by journal or roller bearings, and axial positioning is left to thrust

bearings. Heavy duty applications are usually dominated by journal bearings, while

aerospace applications tend to use ball/roller bearings.

4.2.1. ROLLING BEARINGS

These bearings are differentiated according to

the direction of the main radial or axial loads, as

well as the type of rolling elements which can be

balls or rollers. Balls are used for lower carrying

capacities and higher speeds, while roller

bearings have higher load capacity and lower

speeds.

Aeroderivative gas turbines, as well as jet

engines – where lightweight components are

used due to lower loads, requirement of the

vehicle, smaller size etc – typically use roller

bearings throughout since these do not require

lube oil reservoirs, coolers and pumps associated with journal bearings.

Figure 10. Different

types of rolling

elements: Ball (left),

rolls (center), tapered

rolls (right).

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4.2.2. JOURNAL BEARING

Journal bearings use the principle of Hydrodynamic

film lubrication. A shaft (or journal) spins inside a

stationary sleeve, which has a slightly bigger bore

which is filled up with a lubricant. When the shaft starts

spinning, the lubricant is pushed between the shaft and

the sleeve creating a film, eliminating the direct contact

between the two components, usually made out of

metal.

The application of this bearings is largely used in

heavy machinery due to the high loadcarrying

capacity (radial and axial loads). Therefore, in gas

turbines it is mostly seen in industrial grade energy generation turbines.

For radial loads, it is worth to make a special mention to the Tiltingpad bearing, which is

the most common type in today’s machines. It consists in several pads located around

the shaft, which are floating in order to adapt the best position during operation.

In journal bearings, axial loads can also be handled by thrust bearings. These are used

to resist the unbalanced force in a machine’s working fluid and maintain the rotor in its

position. Tiltingpad thrust bearings are available and, when perfectly aligned, are

notorious to being advantageous compared to taperland bearings.

25


Figure 12. Diagram of a Turbofan showing the bearings supporting the connection between the

different rotating elements of the engine.

The bearings used in turbomachinery is a combination of the described in this section of

the report. The number of bearings is defined by the number of spools of the engine.

Figure 12 shows the placement of bearings along the coaxial shafts of a three spool

engine –particularly of a turbofan engine used in aerospace– where bearings support

the connection between the compressor and turbine shafts at low and high pressure

stages. Since this is considered to be a low loadcarrying engine, the entirety of the

bearings seen in the figure are rolling bearings. In table 8 the direction of the loads in

each section are discussed, which will determine the type of bearing suitable for the

application.

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Section Bearing Interface Type of Load Suggested Bearing

A LPC rotor to IPC Axial Ball Bearing

B LPC shaft drive support Radial Roller Bearing

C HPC rotor to IPC Axial Angular Ball Bearing

D HPC and HPT driveshaft support Radial Roller Bearing

E LPC and LPT driveshaft Radial Roller Bearing Table 8. Bearing selection per section of the engine.

In industrial applications for energy generation, the dimensions of the components is

much larger, as well as the loads that the bearings have to cope with. In this case the

use of journal bearings is more common. However, journal bearings designed to resist

thrust are not able to resist radial loads, therefore wherever an angular ball bearing has

been used in the turbofan

shown above, there will need to

be a combination of a thrust

bearing and a radial journal

bearing.

Figure 13. Combination of a journal and thrust bearing

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5. DISCUSSION

The efficiency of the cycle has been calculated to be of 39.39%. This efficiency leads to

calculate loading factors between 1.32 and 1.16, which are acceptable values for the

Reaction of this scenario.

Even though this seems a rather low efficiency for power generation, this gas turbine

must be coupled to, for example, a steam turbine that uses the high temperature air at

the nozzle. This way, with a combined cycle, the overall efficiency can be increased to

acceptable values and make the most out of the energy entering the system.

All the parameters calculated are within common values of turbine designs, therefore

the preliminary design gathered in this report is accepted to proceed with the next

stages of the design of the turbine.

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6. REFERENCES

Raj, Rishi “ThermoFluid Systems Analysis and Design, Emmerson: Publication of GI Corporation, 2016 Eighth Edition”. Boyce, Meherwan P. “Gas Turbine Engineering Handbook. Amsterdam: Elsevier / Butterworth Heinemann, 2012”.

Web sites

https://link.springer.com/article/10.1007/s1183701620712

http://ae3006uee.blogspot.com/2007/10/

https://www.quora.com/Howisaturbofanengineshaftsupportedinsidetheenginecas

ing

https://www.google.com.ar/patents/US6378293

29

https://link.springer.com/article/10.1007/s11837-016-2071-2

http://ae3006-uee.blogspot.com/2007/10/

https://www.quora.com/How-is-a-turbofan-engine-shaft-supported-inside-the-engine-casing

https://www.quora.com/How-is-a-turbofan-engine-shaft-supported-inside-the-engine-casing

https://www.google.com.ar/patents/US6378293

Documents

DESIGN PROJECT: DESIGN OF A GAS TURBINE