A Quick Method for Full Flange-to-flange Industrial Gas Turbine Analysis Based on Through-flow Modelling

Milan V. Petrovic1, Ahmed Abdel-Rahman and Alexander Wiedermann2

1 University of Belgrade, Faculty of Mechanical Engineering, Laboratory of Thermal Turbomachinery

Kraljice Marije 16, 11120 Belgrade 20, Serbia 2 MAN Diesel & Turbo SE, Steinbrinkstrasse 1, 46145 Oberhausen, Germany

DRAFT: International Journal of Gas Turbine, Propulsion and Power Systems

ABSTRACT In this paper flow field calculations for an entire two shaft in-dustrial gas turbine will be described. This method is based on individual through flow codes for axial compressors and air-cooled expansion turbines developed by the authors which are automati-cally coupled using simple combustion and secondary flow models connecting compressor and turbine flow paths. With this approach the complete quasi 3-D flow field from compressor inlet to turbine exit can be solved simultaneously (flange-to-flange). Details are explained in this paper. The through-flow computation for the analysis of cooled axial multistage turbines is fed by air from the compressor bleeds which are part of the through flow model of the compressor. The through-flow methods are based on a stream function approach and a finite element solution procedure. They include high-fidelity loss and deviation models with improved correlations. Advanced radial mixing and endwall boundary layer models are applied to simulate 3-D flow effects. For air-cooled turbine analysis, various types of cooling air injection were en-compassed: film cooling, trailing edge injection and disc/endwall coolant flow. Compressor and turbine flow path computations were extensively validated individually and published by the authors. Predicted gas turbine operating points of MAN’s MGT-gas turbine will be compared with results of the 3-D Navier- Stokes solver TBLOCK which was run for both compressor and turbines indi-vidually using the boundary conditions derived from the present analysis. The focus is on the comparison of mean data and radial distributions at inlet and outlet stations as well as planes between individual stages and blade rows. They will be compared with measured data at MAN’s gas turbine test rig which were obtained in the turn of a prototype telemetry test campaign. It will be demon-strated that the new method presented is an essential and quick tool for overall gas turbine design and matching of the gas turbine components. NOMENCLATURE

c = Velocity i = specific rothalpy h = specific enthalpy

0h = total enthalpy

Bk = bleed mass flow coefficient

pk = pressure loss coefficient

m = mass flow rate

p Pressure r radial coordinate, radius s specific entropy T temperature z axial coordinate

Subscripts

1 = compressor inlet 2 = compressor outlet 3 = turbine inlet 4 = turbine outlet B = bleed C = coolant CC= combustion chamber cool = cooling F= fuel g = gas mix = mixed flow r.mix= radial mixing u = circumferential component z = axial component

INTRODUCTION

The design of a new gas turbine starts with definition of the power output and main parameters of the cycles. Then, based on cycle calculation, experience and the expected level of technology that will be applied, the mass flow rate and design parameters for the compressor, combustion chamber and turbine are set. Subse-quently, the design process of each machine component runs more or less individually in separate design departments. Compressor and turbine departments apply meanline analyses to lay the merid-ional gas path configuration, number of stages and stage work distribution. 2-D codes are used for blade geometry definition, parameter optimization and performance calculation. Finally, 3-D codes are applied for further improvement of the design by mini-mization of losses, extending the operating range and improvement of part load behaviour. In spite of the great progress in numerical methods and computer capacity, 3-D simulations of multistage machines are still limited by long execution times and, therefore, only a limited number of calculations can be performed [1,2].

As the result of all these efforts, the internal efficiency of each individual component, namely compressor, turbine and combustion chamber, has reached very high levels. The next step in the design process is simulation of the performance of the complete machine, matching of the turbine and the compressor, and simulation of off-design operation.

Since here the joint operation of the turbine, combustion chamber and compressor should be simulated, usually only simple

International Journal of Gas Turbine, Propulsion and Power Systems May 2016, Volume 8, Number 1

Presented at International Gas Turbine Congress 2015 Tokyo November 15-20, Tokyo, Japan Manuscript Received on Janualy 19, 2016 Review Completed on April 18, 2016

Copyright © 2016 Gas Turbine Society of Japan

9

1-D codes are used for this purpose [3]. They are quick, but the results are not sufficiently accurate and reliable and suffer from a lack of relevant information. For this reason, such a type of simu-lation cannot be used for design parameter optimization.

Results of a fully coupled system for industrial gas turbine analysis are presented in this paper. The coupled method has been applied to the operation analysis and performance prediction of a newly developed industrial gas turbine and was presented recently [4]. This method has been extended by an automated coupling procedure of the program suite using FORTRAN programmes and Linux shells developed at MAN. The main focus is on the applica-tion to two-shaft gas turbines, where the gas generator power bal-ance must be solved by an iterative approach. Predicted gas turbine operating points are compared with experimental test data and results of the 3-D Navier-Stokes solver TBLOCK which was run independently for both compressor and expansion turbines using the boundary conditions of our new through-flow flange-to-flange solver. The results demonstrate the capability of the method to predict the joint operation of compressor, combustor, HP and power turbines properly DESCRIPTION OF THE METHODS

The quasi- 3D flange-to-flange method is based on individual through-flow methods for axial compressors and air-cooled gas turbines, developed by the authors, which are coupled using simple combustion and cooling flow models connecting compressor and turbine flow paths.

The through-flow method applied here was developed using theory based primarily on the analyses of Hirsch and Deconinck [5], Denton [6], Petrovic [7] and Petrovic and Riess [8], but it includes many new details and extensions. The flow is assumed to be steady state, adiabatic and axisymmetric, by which a two-dimensional description is achieved. The 2-D calculation is carried out within the meridional hub-to-shroud surface, at which the momentum equation is projected. Body forces are introduced to replace turbine blades, and flow loss effects are approximated by a friction force. The method is based on a stream function approach and a finite element solution procedure. Through-flow Method for Axial Compressors (ACFlow)

The through-flow method for flow calculation in multistage axial compressors was developed by the authors and presented in detail in previous papers [9–11].

To achieve high accuracy and robustness of the method, a new loss and deviation model for compressor cascades was developed, tested and applied [9,11]. Taking into account the decisive influ-ence of flow in the inlet guide vane (IGV) on the compressor per-formance, extensive investigations were undertaken with the aim of developing a new systematic methodology for the prediction of IGV performance [10]. Reference loss and deviation correlations, which include the main geometry parameters and Мach number influence, are supplemented by correlations for additional losses and deviations due to stagger-angle adjustments for off-design operation. Sets of correlations to calculate the cascade exit flow angle deviation and total pressure loss were created in such a way as to include a wide range of geometry and operating condition parameters and their influences. Also, the loss and deviation mod-els include shrouded clearance, tip clearance and secondary flow losses with the spanwise distribution and their mutual interaction [11].

The endwall boundary layer model is coupled with the main flow calculation by an additional annulus blockage factor. The spanwise-averaged meridional velocity is used as the model input. The endwall boundary layer equation system is integrated over the

local portion of the endwall length and the problem is solved further through an iterative procedure. The model includes the influence of bleed suction and also calculation of entropy production in the endwall boundary layer.

A significant part of the method is a model for full blade stall and compressor surge limit prediction based on geometry and flow parameters that allows the definition of appropriate design param-eters taking into account the required operating range.

For given inlet conditions (mass flow rate, inlet pressure and temperature of the air) and flow path geometry, the results of cal-culation are as follows:

overall compressor performance, i.e. efficiency and pressure ratio for different speeds and guide vane adjustments;

thermodynamic properties and flow field (velocities, stream function, Mach number) at the outlet and at each point of the compressor meridional plane;

blade row and stage parameters such as spanwise distribu-tion of losses, efficiency, reaction, load and flow parameters and stall detection parameters.

The method is applicable for subsonic and supersonic flow in compressors for a very wide range of operation: from stall and surge to the choke limit.

The method was tested, calibrated and validated, comparing the numerical results with experimental data for a large number of test cases [10]. These test cases include compressors with very different configurations and operating ranges. Figure 1 shows a validation example for a typical industrial 11-stage gas turbine compressor [10,11].

Fig. 1 Overall performance of MAN 11-stage compressor, ACFlow results and rig experiments

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Through-flow Method for Air-cooled Gas Turbines (GTFlow)

The development of the through-flow method for flow analysis in axial flow gas and steam turbines was described in previous papers [8,12,13]. Calibration and validation of the method were performed by detailed calculation of the flow in a number of un-cooled turbines.

In the case of air-cooled turbines for determination of rothalpy, i.e. total enthalpy, angular momentum and specific entropy along a streamline, the following equations are available:

.

.

.

r mix cool

u u u ucascade r mix cool

loss r mix cool

i i irc rc rc rc

s s s s

(1) (2) (3)

The terms with subscripts r.mix and cool denote the influence of radial mixing and cooling effects on the main-stream flow. To encompass these effects, an appropriate spanwise mixing model and cooling model was developed and implemented.

A new improved and calibrated loss and deviation model is applied to calculate the entropy variation, losss , and the circum-

ferential component of the flow velocity, urc . 3D flow effects were modelled by newly developed models for the radial distribu-tion of losses and spanwise mixing [8].

For air-cooled turbine analysis, various types of cooling air in-jection were introduced: film cooling, trailing edge injection and disc/endwall coolant flow.

Fig. 2 Definition of cooling model for stator (film cooling and TE

coolant ejection) and disc/casting cooling

Figure 2 shows schematically the application of the mixing model for the first stator with one row of film cooling holes down-stream of the leading edge and ejection at the trailing edge. Disc/endwall cooling also exists. Due to coolant injection, the mass flow rate, specific total enthalpy and specific entropy of the main stream are changed. For the mixing process, the following balance equations are conserved:

mix g cm m m

(4)

0 0 0g cmix g c

mix mix

m mh h hm m

(5)

mix gs s s

(6)

where 0h and s are the mass-averaged total enthalpy and entro-py in the main stream. Subscript g is used to denote gas, c to denote coolant and mix for flow after mixing. The complete air cooling model was presented in a previous paper [12]. However, it is clear that accurate information about coolant properties is nec-essary for a successful calculation of flow in the turbine. This information can be obtained only by accurate calculation of flow in the compressor for each single operating point.

Turbine flow path computations were extensively validated by comparison of the numerical data with experimental data for a typical modern industrial gas turbine [12]. For example, the ex-pansion line in the turbine is presented in Figure 3.

Fig. 3 Enthalpy-entropy diagram of the expansion line in MAN

turbine with air cooling

Process in Combustion Chamber (CCPro)

The process in the combustion chamber is described here by a simple 1D combustion calculation. The operation performance and design parameters, such as efficiency, relative pressure drop and profiles of temperature and pressure at the combustion chamber outlet should be known. For a given mass flow of air, pressure and the temperature of air at the compressor outlet and the required temperature of gases at the turbine inlet, this module calculates:

o the composition of the combustion gases, o fuel consumption and mass flow rate of combustion

gases, o pressure and other gas properties at the chamber outlet, o spanwise distribution of pressure and temperature of

combustion gases at the turbine inlet is taken from the experiments.

To calculate the properties of air and combustion gases, the correlations of Baehr and Diederichsen [14] are applied. FULLY COUPLED SOLUTION METHOD FOR OVERALL GAS TURBINE SIMULATION

The fully-coupled method is based on the three individual through-flow codes described above. These are linked by the main

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flow and the cooling flow paths (Fig. 4). Flow is extracted from the compressor main flow at bleed ports after stages 4, 8 and 11 and bypasses the combustion chamber. These bleed flow rate shares cover 1. the amount of cooling flow rates that re-enter the main flow

path at the coolant flow ejection stations in the turbine (Fig. 5), 2. leakage flow bypassing the turbine without being injected

back into the flow path behind the last turbine stage exit, and 3. discharged flow at the bleed ports for start-up of the gas tur-

bine and operating conditions near to idle.

Fig. 4 Structure of the method

The share of the flow of compressor extraction stations 4, 8, and 11 for turbine cooling is bifurcated and re-ejected into the turbine at side walls and blade leading edges and through film cooling holes as described above. The scheme of the overall cool-ing flow system is shown in Figure 5 with 3 compressor extraction and 17 turbine injection stations. Temperature and pressure differ-ences between compressor extraction and turbine injection stations are determined by an external hydraulic network prediction system for the cooling air flow path system of the turbine which is fed by the compressor bleed flow parameters.

The individual codes are interconnected, controlled by an outer system of Fortran programs and Linux shells in order to transfer data between the codes, establish the overall conservation of mass, and find the solution for given rotational speeds and turbine exit parameters. An iteration procedure is necessary, which differs for single- and twin-shaft gas turbines, and is illustrated in Fig. 6. Single-shaft Gas Turbine Solution Procedure

Though single-shaft gas turbine analysis shall not be discussed in this paper we will shortly explain the procedure for complete-ness. For a single-shaft gas turbine, the procedure is straightforward because the turbine shaft power is given by difference between the overall power generated by the turbine and the sum of compressor consumption, pumping work in the cooling flow system and me-chanical losses. First, the compressor speed line is predicted with ACFlow for a specified rotational speed. In the next step, an initial guess is made for a specific compressor operating point and the turbine inlet temperature. The resulting fuel mass flow is given by CCPro and transfers inlet data to the turbine through-flow solver GTFlow. The turbine inlet temperature is adjusted by an inner iteration loop to meet the prescribed turbine discharge pressure. Finally, a second iteration loop is run to adjust the compressor working point along the compressor characteristic at constant VGV

setting in order to obtain the desired gas turbine operating point, which is determined by an expected value of the turbine exit tem-perature. Twin-shaft Gas Turbine Solution Procedure

The solution procedure is more complex in this case because the gas generator and the power turbine are operating on different shafts. The gas generator system consisting of the compressor and high-pressure turbine must be balanced out and, in addition to the procedure for the single-shaft gas turbine, the gas generator speed must be adjusted such that the power generated by the high- pres-sure turbine is fully consumed by the compressor and the shares corresponding to cooling flow pumping work and mechanical losses. The target of the second, outer iteration loop is a common rotational speed of compressor and turbine to determine the oper-ating point of the gas turbine. As the power turbine works inde-pendently at different speeds, this procedure has to be repeated for each power turbine rotational speed.

Fig. 5 Cooling air flow distribution

Fig. 6 Flow-chart of solution procedure 3-D VISCOUS SOLVER TBLOCK

For the components compressor and turbine 3-D viscous com-putations were carried out with a customized version of the mul-

guess T0

CCPro

GTFlow Revise T0

Select next

compressor

operating point

(Pout – ptarg) < ɛ ?

Balance of GG-

power?

ACFLOW

End of single-shaft

solution procedure

End of twin-shaft solution procedure

Specify ptarg at turbine exit

pout = f(T0)

no

no

yes

yes

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ti-block Navier-Stokes solver TBLOCK, which had been originally developed and extensively validated for turbine flow analysis by Denton [15,16]. The code had been interfaced with MAN’s design and analysis environment and calibrated with numerous benchmark test data of technical interest for gas turbine design. This solver is executed both for steady-state and unsteady analysis. For the latter case, TBLOCK was recently verified and included in a process chain to handle with blade flutter [17]. The parallelized MPI ver-sion was further developed by MAN and is currently used as an efficient working horse for multistage turbomachinery analysis and components of the secondary flow system.

TBLOCK has been validated for axial compressor flow analy-sis using data of the open literature as well as test data obtained at rig tests of MAN’s 11-stage compressor [18]. The performance data and the blade pressure contours for the front stages are shown in Figure 7.

Fig. 7: TBLOCK analysis of MAN’s 11-stage compressor [18]

Because of its multi-block structured mesh topology it is easily possible to include the stator and rotor shroud geometries and cavities in the turbine main flow field analysis to obtain insight into the interaction of main and secondary flows. Cooling flow ejection

into the main flow can be simulated by defining special source and sink patches along each arbitrary surface, Figure 8 shows averaged relative Mach number contours in the meridional S2-plane of the HP- and power turbines of the MGT-gas turbine which are ana-lyzed simultaneously, examples for the interaction of shroud and cavity flows with the main flow are shown in Figure 9. Using 25 processors, fully converged solutions can be obtained overnight with over 11 million grid points.

Fig. 8 TBLOCK analysis of HP- and power turbines of MAN’s MGT- gas turbine

Fig. 9: Details of TBLOCK analysis of HP- and power turbines of

MAN’s MGT gas turbine METHOD APPLICATION

In this section results of the present analysis for MAN’s two shaft industrial gas turbine of the MGT-series are discussed and verified against experiments that were performed at the gas turbine test site of MAN’s Oberhausen plant. The overall analysis was conducted with the flange-to-flange through flow solver. In addi-tion, full 3-D Navier-Stokes analyses were run using the results of the flange-to-flange solver for each component, compressor and turbine, individually. Averaged results and radial profiles of essen-tial flow parameters have been compared with each other and experiments. Flange-to-flange Results of Through-flow Suite

The twin-shaft gas turbine is optimized for mechanical drive applications and features a power turbine that is designed for a wide operating range. Both single- and twin-shaft MGT series gas turbines have compressor, combustor and stages 1 and 2 of the high-pressure turbine in common. Instead of the third stage of the single-shaft engine, the high-pressure turbine is followed by a rear-stage structure to house the gas generator bearing and a free two-stage power turbine designed for a wide operating range. This engine has been tested extensively at the Oberhausen gas turbine test rig and successfully passed its long-term testing in a CHP application at the launch customer’s site.

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Figure 10 shows the distribution of the pressure and the temperature in the meridional plane of the twin-shaft turbine calculated with our flange-to-flange through-flow solver suite.

Fig. 10 Pressure and temperature contours in the MAN twin-shaft

gas turbine

Fig. 11 Cycle simulation – twin-shaft turbine

Figure 11 displays a typical h–s diagram obtained with our

approach. The slope of the turbine expansion line clearly shows the influence of the cooling flow.

As explained in the previous section, the gas turbine operating line is given by balancing the turbine work with the compressor consumption and the work losses associated with mechanical losses and cooling flow pumping work. As an example of intermediate results of the iterative approach, turbine and compressor power are plotted in Figure 12 for various rotational speeds against compressor inlet mass flow. Each operating point of the compressor characteristic corresponds to a converged solution for the coupled CCPro and GTFlow computations. The unique solution is indicated by the cross-over point of the graphs for compressor and turbine powers as a result of the outer iteration loop.

A comparison of predicted and measured gas turbine operating conditions (reduced to ISO inlet conditions) obtained during one day of our test campaign is shown in Figure 13. Here, the operating regime ranges from less than 30% part load up to full load. The

cooling air flow pumping work is assumed to vary as function of the rotational speed of the gas generator, and mechanical losses have been assumed constant. Good agreement can be seen; however, the measured data of the operating line shown in Figure 13(a) are shifted slightly towards lower values. On the other hand, power turbine inlet pressure (c), total temperatures at HPT inlet, LPT inlet and turbine exit (b) and shaft power vs. mass flux are well predicted over the entire operating range.

Fig. 12 Computation of the gas turbine operating line for twin-shaft engine

(a) Predicted and measured total pressures at compressor exit

(1) and LPT inlet stations (2)

(b) Predicted and measured total temperatures at HP-inlet (1),

LP-inlet (2) and turbine exit (3) stations

s [kJ/kg]

h [k

J/kg

]

1

42

3

Mass flow rateSh

aft p

ower

94,4%95%

96%97%

98%99%

100%

101%102%

Compressor HP - Turbine

1 kg/s

1MW

Tota

l pre

ssur

e ra

tio

Mass flux

Predicted

Experiments

Tota

l tem

pera

ture

Mass flux

Predicted

Experiments

(1)

(2)

(3)

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(c) Predicted and measured shaft powers

Fig. 13 Twin-shaft turbine – operating line

Computation of Gas Turbine Operating Points Using Through-flow System and 3-D CFD TBLOCK

In this chapter our proposed procedure is discussed to compute detailled data of the 3-D flow field with the multi-stage solver TBLOCK based on the results of the through-flow solver suite which describes the coupling of the gas turbine components. Because turbine and compressor powers must be carefully balanced out to maintain the desired gas generator speed the matching point must be determined iteratively as shown in Figure 12. This means that the set of boundary conditions obtained with our new flange-to-flange approach is the most realistic assumption for the calculation of the operating conditions of each compressor, HP- and power turbines individually with our 3-D Navier-Stokes solver TBLOCK. Though the through-flow code itself offers many valuable information about stage-matching, radial flow profiles between the blade rows and performance data, results of 3-D flow analysis are required to evaluate more flow details as effects of loading on blade profile pressure distributions, secondary flow and local flow separation and so on. Picking out one individual operating point, results of this approach will be presented, and mean values as well as circumferentially averaged flow quantities of the 3-D solution will be compared with the results of the through-flow solver and experiments, where available.

(a) Mean-averaged pressure distributions along the com-

pressor

(b) Mean-averaged pressure distributions along HP- and

power turbines Fig. 14 Gas turbine overall total and static pressure lines vs. length

Figure 14 shows the averaged pressure distribution along the axial direction for both compressor (a) and turbines (b). A good comparison can be found between the slopes of both solvers and pressure readings taken along the outer casing. In contrast with the GTFlow where diffuser losses were estimated by correlation, the compressor and turbine exhaust diffusers are part of the 3-D flow analysis. This is the reason for minor differences in the overall compressor pressure ratio prediction. Total temperature profiles at the leading edge planes of the 1st nozzle guide vane and stator 2 are shown in Figures 15 and 16. The radial temperature distribution at the combustor exit was estimated based on 5 radially located total temperature probes applied to the leading edge of 6 selected nozzle guide vanes to resolve the lateral variations due to the can combustors of the gas turbine. The slopes shown in Figure 15 represent an estimation of the circumferentially averaged temperature distributions which were used to prescribe the radial total temperature profiles for both the through-flow and 3-D Navier-Stokes analyses. Small differences between Throgh-Flow and CFD data are due to the averaging procedures applied.

Fig. 15 Nozzle guide vane inlet temperature profile for the turbine

flow field analysis

At the inlet station of HP Vane 2 the experiments offer a more homogenous distribution than the 3-D results obtained with the multi-stage solver TBLOCK, Fig. 16. It is obvious that the 3-D results are characterized by the occurrence of strong secondary flow interacting with the cooling flow entering the flow passage

GT

Shaf

t Pow

er

Mass flux

Predicted

Experiments

Pres

sure

Axial length

TBLOCK - Ptot

TBLOCK-Pstat

GTFlow Ptot

GTFlow- Pstat

Exp

Pres

sure

Axial length

TBLOCK - Ptot

TBLOCK-Pstat

GTFlow-Ptot

GTFlow-Pstat

Exp

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

Frac

tion

of s

pan

Total temperature

TBLOCK

Experiment

GTFLOW

50 K

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downstream of the 1st stage. A more detailed look at the predicted streamline pattern along the suction side shows strong secondary flow traces due to the passage vortex flow near ID and tip vortex flow moving down from OD. Both vortices do not merge near mid span in contrast with the observed more uniform measured slope. It is remarkable to note that the results of the through flow solver follow the slope of the experiments very closely.

Fig. 16 Total temperature profile at the leading edge plane of 2nd

stage stator

Fig. 17 Total temperature (above) and pressure (below) in the

interstage diffuser between HP and power turbines

Total temperature and pressure profiles between the high-pressure and power turbine are shown in Figure 17 and com-pared with averaged measured data of 6 rakes distributed in cir-cumferential direction, with 7 radial probe positions each. The agreement is reasonably good and demonstrates that the matching of HP and power turbines has been captured well with our analysis.

Figure 18 shows the corresponding total pressure and temperature profiles at the power turbine exit station at the diffuer inlet. The variation band of all results is almost within a 10°K range. The outlet pressure distribution is obtained in close agreement with the measured slope with the 3-D Navier Stokes solver TBLOCK. It must be mentioned here that the 3-D viscous flow analysis includes the diffuser flow as well. On the other hand, radial equilibrium was specified at the rear rotor exit for the through flow analysis which is reflected in the predicted slope of the pressure line.

Fig. 18 Total temperature (above) and pressure (below) at the turbine exit plane

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

Frac

tion

of s

pan

Total temperature

TBLOCK

GTFLOW

Exp

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

Frac

tion

of s

pan

Total Pressure

TBLOCK

GTFLOW

Exp

0

0.1

0.2

0.3

0.4

0.5

0.6

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1

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of s

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TBLOCK

GTFlow

Exp

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0.9

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Frac

tion

of s

pan

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TBLOCK

GTFlow

Exp

50 K

50 K

50 K

0.5 bar

0.5 bar

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(a) NGV exit plane

(b) HP Rotor 1 exit plane (relative frame)

(c) HP Stator 2 exit plane

(d) HP Rotor 2 exit plane (relative frame)

(e) LP Stator 1 exit plane

(f) LP Rotor 1 exit plane (relative frame)

(g) LP Stator 2 exit plane

(h) LP Rotor 2 exit plane (relative frame)

5°

0.1 5°

5°

5°

5°

10°

5°

10°

0.1 0.1

0.1

0.1

0.1

0.1

0.1

Fig. 19 Radial flow angle and Mach number distributions at each blade and vane exit planes of two-shaft gas turbine

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Figures 19 (a) – (h) show results of the absolute and relative exit flow angles and Mach numbers at the individual blade row exit planes. These comparisons of through-flow and 3-D CFD results of TBLOCK indicate that both solvers give very similar characteris-tics for the work distribution in each blade and vane row of the HP and power turbines. Differences are more pronounced in the front stages where secondary flow effects and their interaction with injected cooling flow between the platforms are extremely strong. Except the regions near the side walls, the predicted slopes are in very good agreement and are nearly identical at the exit stations of some of the blade rows.

CONCLUSION

A model for the simulation of complete gas turbine aerody-namics based on through-flow analysis has been presented. It con-sists of a suite of three individual solvers for compressor, turbine and one-dimensional combustor analysis, which have been fully coupled and solved iteratively to meet the real flange-to-flange boundary conditions of gas turbine operation. The model has been applied to the simulation of two-shaft gas turbine flow field. As an example MAN gas turbines of the MGT series were investigated where experimental data were available. The comparison of pre-dicted and measured data shows reasonable agreement and suggests that slopes and trends can be well predicted with this approach. Further refinements are necessary to evaluate mechanical losses and take into account the effect of additional loss mechanisms in the cooling flow system.

Overall, in addition to its relevance for the design process, this approach is very promising for evaluating gas turbine behavior based on field data. The effect of changes and leakages in the cooling flow system due to degradation can be traced back with this model, and also failures in VGV settings. Malfunctions of the gas turbine that might be a precursor of severe damage could be de-tected with a well-calibrated system. The presented coupled meth-od is useful for diagnosis and supports the root-cause analysis of misbehaving field engines. REFERENCES [1] Denton, J. D., Dawes, W. N.,1999, “Computational Fluid

Dynamics for Turbomachinery Design“, Proc Instn Mech En-grs, Vol 213 Part C, pp 107/124

[2] Jennions, I. K., 1994, “Elements of a modern turbomachinery design system,” in: Turbomachinery Design using CFD, AGARD-LS-195

[3] Dubitsky, O., Wiedermann, A., Nakano, T., Perera, J., 2003, “The reduced Order Through-Flow Modeling of Axial Tur-bomachinery”. Proc. Int. Gas Turbine Congress 2003, Tokyo, Paper TS-052.

[4] Petrovic, M. V., Wiedermann, A., 2015, “Fully-coupled through flow method for industrial gas turbine analysis,” ASME Paper No. GT2015-42111,

[5] Hirsch, Ch., Deconinck, H., 1985, “Through Flow Models for Turbomachines: Stream Surface and Passage Averaged Repre-sentations,” in: Thermodynamics and Fluid Mechanics of Tur-bomachinery, Vol. I, ed. Ucer, A. S., Stow, P., Hirsch, Ch., Martin Nijhoff Publishers, Dordrecht.

[6] Denton, J.D., 1978, “Throughflow Calculation for Transonic Axial Flow Turbines,” ASME, J. Eng. Gas Turbines Power, 100(2), pp. 212–218.

[7] Petrovic, M. V., 1995, “Meridional Flow Calculation in Multi-stage Axial Turbines under Design and Off-Design Condi-tions”, VDI Series 7, No. 280, VDI-Verlag, Düsseldorf, Ger-many (in German).

[8] Petrovic, M. V., Riess, W., 1997, “Off-Design Flow Analysis and Performance Prediction of Axial Turbines,” ASME Paper No. 97-GT-55.

[9] Petrovic, M. V., Wiedermann, A., Banjac, M., 2009, “Devel-opment and Validation of a New Universal Through-Flow Method for Axial Compressors,” ASME Paper No. GT2009-59938.

[10] Banjac, M., Petrovic, M. V., Wiedermann, A., 2014, “A New Loss and Deviation Model for Axial Compressor Inlet Guide Vanes, “ASME. J. Turbomach., 136(7), pp. 071011.

[11] Banjac, M. B, Petrovic, M. V., Wiedermann, A., 2014, “Secondary Flows, Endwall Effects and Stall Detection in Axial Compressor Design, “ASME Paper No. GT2014-95020.

[12] Petrovic, M. V., Wiedermann, A., 2013, “Through-Flow Analysis of Air-Cooled Gas Turbines,” ASME J. Turbomach., 135(6). pp. 061019.

[13] Petrovic, M. V., Riess, W., 1997, “Off-Design Flow Analysis of LP Steam Turbines,” Proc Instn Mech Engrs, Part A, 211(A3), pp. 215–223.

[14] Baehr, H. D., Diederichsen, C., 1988, “Berechnungsgleichun- gen für Enthalpie und Entropie der Komponenten von Luft und Verbrennungsgasen” (in German), BWK, Vol. 40, pp 30-33.

[15] Denton, J. D., 2005, “Multiblock CFD – Solver TBLOCK”, Cambridge University, UK

[16] Rosic, B., Denton, J. D., Curtis, E. M., 2007, “The Influence of Shroud and Cavity Geometry on Turbine Performance,” ASME GT2007-27769 and -27770 (part 1 + 2)

[17] Micallef, D., Wittek, D., Wiedermann, A., Mailach, R, 2014, „An Efficient Workflow for Accurate Flutter Stability Anal-yses and Application to a State-of-the-art Compressor Rotor,” ASME-Paper GT2014-25646, Düsseldorf

[18] Wiedermann, A., Frank, D., Orth, U., Beukenberg, M., 2011, “Computational and Experimental Analysis of an Industrial Gas Turbine Compressor,” ASME-Paper GT2011-46336, Vancouver

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