Marco Grosso

Federico Sansoni

Alessandro Sorce
email: alessandro.sorce@unige.it

Margherita Monti

Fabio Pascucci

Roberto Razzoli

Influence of Start-up Management on the Residual Life of a Large Steam Turbine Shaft

Keywords:

établissements d'enseignement et de recherche français ou étrangers, des laboratoires publics ou privés.

INTRODUCTION

The progressive liberalization of electrical energy and natural gas in Europe introduced a growing competition between energy producers, making fundamental the ability to optimize management strategies and control of the plants to reduce production and service costs. This process must also take into account strict environmental regulations on pollutants emission.

The ability of the operators in optimally managing the facilities is essential to be competitive in the liberalized market. The variability of the fuel cost, the instability of the electricity demand and its price, the continuous increase of power production from renewable sources, create complex problems to operators and force them to develop operational models for improving management systems in terms of efficiency, flexibility and reliability.

All these aspects characterize the modern control strategies of the power generation units; in fact, a management method which can operate the plant in a highly discontinuous and irregular manner is crucial for:  Meeting the users' demand and the severe regulations on emissions of macro pollutant (mainly CO and NOx)  Safeguarding the stability of the electrical grid  Maximizing profits

The power plant should be able to produce more power during peak hours, where the gain is higher, whilst reduce the power supplied to the grid or turns the production unit off in the hours of low demand. Therefore, each unit must be able to operate both at high loads and to remain off or at the minimum during the night and the weekend, but also to restart in a short time. The start-up time is strictly related to the initial status of the power plant which depends basically from the standstill time. After 8 hours of standstill, an hot start up can be performed in less than 60 minutes; a warm start-up, typically after the weekend standstill (8÷48 h) takes about 80÷120 minutes, while the cold start up requests from 2 to 3 hours [START_REF] Kelhofer | Combined-Cycle Gas & Steam Turbine Power Plants[END_REF].

Gas turbines and combined cycle are more flexible than other types of thermal power plants (fuel oil or coal), and are often called to operate in a discontinuous manner characterized by frequent and rapid variations of the produced power [START_REF] Balling | Flexible future for combined cycle. Modern power systems[END_REF]. Nevertheless, the new requirements push all plants, even those initially designed to produce as base load, to change the type of operation or switch to a more flexible management for maximizing the economic return and safeguarding the electrical grid stability [START_REF] Greis | Flexible and Economical Operation of Power plant: 25 Years of Expertise[END_REF].

A flexible management ensures higher profits in the short term, but reduces the life of critical components of the system because of the damage caused by thermo-mechanics fatigue phenomena that arise during load changes, start-ups and shutdowns. The parts subjected to high temperature as the gas turbine and the steam components suffer significant cyclical variations in temperature and pressure parameters that induce stresses on the material, thus damaging the components [START_REF] Greis | Flexible and Economical Operation of Power plant: 25 Years of Expertise[END_REF]. This means loss of profit due to frequent failures or the need of ordinary and extraordinary maintenance.

Of course, the most stressful maneuvers are the plant start-ups and shut-downs, which affect the residual life of the components and the reliability of the whole system. In order to have the best performance during the entire life of the plant, the operator must be aware of the relation between the residual life of the component and the rate of load change to provide guidance on the long-term consequences. Gas Turbine (GT) manufacturers take into account for the additional life consumption during start up by summing the Equivalent Operative Hour to the "fired" Operative Hour or setting a maximum number of allowable start-ups between two maintenances [START_REF] Janawitz | Heavy-Duty Gas Turbine Operating and Maintenance Considerations[END_REF]. Similar counters were introduced for the Steam Turbine (ST), discriminating between the different kind of start-up (hot, warm, cold) [START_REF] Kelhofer | Combined-Cycle Gas & Steam Turbine Power Plants[END_REF], but they can miss part of the whole picture. In fact, the life consumption due to the start-up maneuver is not just related to the initial conditions (initial rotor temperature), but also to the way the temperature ramp is performed (i.e. temperature mismatch and temperature gradient). A simple way to visualize this effect is represented by the Cyclic Life Expenditure (CLE) curves, (e.g., see Figure 6.13 in Viswanathan [START_REF] Viswanathan | Damage Mechanisms and Life Assessment of High-Temperature Components[END_REF]) often given by the manufacturer. Those curves represent the percent life expended per cycle as a function of the rate of steam temperature change, and magnitude of steam temperature change and can be used operatively in the start-up optimization process.

Several systems were installed by Original Equipment Manufacturers (OEM) of Steam Turbines or other companies to keep under control the stress of the steam turbine [START_REF] Greis | Flexible and Economical Operation of Power plant: 25 Years of Expertise[END_REF][START_REF] Calzolari | la flessibilità operativa dei cicli combinati coniugata all'efficienza energetica[END_REF][START_REF] Electric | OpFlex TM Steam Turbine Agility Advanced Start-Up for Combined Cycle Plants[END_REF]. For example the Siemens Turbine Stress Controller allows different start-up modes which can be preselected by the operator depending on the grid requirements [START_REF] Kelhofer | Combined-Cycle Gas & Steam Turbine Power Plants[END_REF][START_REF] Greis | Flexible and Economical Operation of Power plant: 25 Years of Expertise[END_REF]. A similar approach is proposed by Ansaldo Energia: the temperature change is monitored with the so called Rotor Stress Evaluator (RSE) and the stress is maintained under some fixed thresholds limiting the GT and ST load rate [START_REF] Calzolari | la flessibilità operativa dei cicli combinati coniugata all'efficienza energetica[END_REF]. The topic is not new, the Electric Power Research Institute (EPRI), developed a big study in 1994 [START_REF]task: Thermal and Stress Analysis of Rotor[END_REF], but is becoming more crucial, as shown by two recent patents developed by O&M [START_REF] Ehrsam | Startup method for large steam turbines[END_REF][START_REF] Sathyanarayana | Method of starting a steam turbine[END_REF] In this paper the focus is on the evaluation of the stresses on the rotor shaft of a steam turbine of a combined cycle gas-steam power plant for the production of electrical power. It is inside the Tirreno Power thermal plant located in Vado Ligure, Italy: the power plant consists of a combined cycle unit (800 MW) and two coal-fired units (330 MW each).

The new combined cycle module started in 2007 and was created to replace an old unit fueled by coal and fuel oil. This choice allowed a significant increase in the efficiency of the system, not to mention that the steam turbine, its auxiliary, the condenser and the generator have been revised and updated to be reused in the new facility. This unit uses two multi-shaft gas turbines (264 MW each), fed with natural gas, and its exhaust goes into a steam turbine (273 MW). Both the gas turbine and the steam turbine were produced by Ansaldo Energia.

TIRRENO POWER'S PLANT MANAGEMENT AND ROTOR STRESS EVALUATOR

The control system of the steam turbine plant optimizes the time of start-up of the groups in the combined cycle. The Rotor Stress Evaluator (RSE) is the system which limits the gradient of ST and GT load in order to reduce the fatigue on the components. It protects the rotor of the machine from exceeding the allowable thermal stress, and therefore limits the thermal stress on the rotor of the steam turbine in two ways:  Preventing the inlet steam temperature mismatch with respect to the rotor surface  Acting on the loading ramps

Figure 1. Rotor Stress Evaluator Calculation Scheme

The Rotor Stress Evaluator is based on an algorithm that processes the temperature of the rotor surface in the section of admission of high and medium pressure steam, an example of probe installation is provided in [START_REF] Greis | Flexible and Economical Operation of Power plant: 25 Years of Expertise[END_REF]. This temperature is considered equal to the measured stator case internal surface temperature, according to the hypothesis that the coefficients of heat exchange of the steam with the metal of the case and the rotor are quite similar (actually the temperature of the rotor is underestimated of about 40°C due to the thermal losses to the ambient).

Then, the average and the center temperature of the rotor is estimated solving -in a simplified manner-the differential equation of heat transfer in a cylindrical body, through an algorithm based on a truncated series expansion.

The stresses are then calculated on the basis of the temperature differences and the permissible stress is calculated as function of the surface temperature (prudential choice).

Starting Probe (SP)

Rotor Temperature Distribution ΔT=f(SP)

Stress Calculation σ=f(ΔT) ST Relative Comparative Stress (σ/G) Permissible Stress G=f(SP)
Run-up, load margins for turbine control A parameter called "margin of load" is evaluated in the rotor region (surface or center) having the slightest difference between calculated stress and permissible stress. The margin of load conventionally produces limitations just when a threshold is overcome: a proportional action is triggered to reduce the load gradient. If the load reduction does not reduce the mechanical load or the calculated stress itself overcomes some security threshold, the RSE trips (i.e. shuts down) the turbine to avoid severe consequences for the rotor.

THERMAL DISTRIBUTION IN THE ROTOR

The first step leads to verify the truncated series expansion results, with respect to analytical calculation in order to use this specific part of the RSE algorithm to estimate the time evolution of rotor's temperature due to different rates of load change. In fact, this is the basis in creating the CLE curves.

So, the RSE program response is compared with the results of analytical calculations for simple input laws. The known analytical models are, in fact, limited by the need to have well-defined boundary conditions.

The geometrical dimensions and physical characteristics used in the calculations are the following ones:

 Rotor diameter, HP stage (first row of blades): 725 mm  Rotor diameter, IP stage (first row of blades): 810 mm

The rotor hasn't a central bore and is a welded disc type. The material is a special alloy for large turbines and generators, having the composition defined as DIN 30CrMoNiV5-11, which mechanical characteristics of can be found in [START_REF] Koenig | Einfluß moderner Verfahren der Stahlerschmelzung auf das Zeitstandverhalten dickwandiger Bauteile aus 1%-CrMoV-Stählen[END_REF].

Therefore, the results are analyzed for the case of the rotor which undergoes a step change in its surface temperature, and a step change in the conditions of the steam lapping it; a onedimensional model of heat exchange in the radial direction is adopted, which allows to considerably simplify the calculations without incurring in large errors.

The first comparison is made with a step change in the surface temperature, see Figure 2. The second one is obtained by a rising temperature of the rotor surface, generated by steam (in constant conditions) entering the machine, see Figure 3. The trend calculated with the analytical model is faithfully reproduced by the RSE. The maximum temperature difference results in the center of the rotor and is about 8° C; moreover, the trend of the average temperature has been calculated correctly, as well. So, as the RSE offers results almost identical to those of the analytical model it was decided to use the RSE series expansion to calculate the temperature distribution, and its values for the calculation of thermo-elastic stresses during transients.

STRESS EVALUATION

The stresses in the rotor are produced by temperature gradients in the material, by the centrifugal force, by the pressure. They can be evaluated separately and combined summing the terms acting through the same axis and then converting the triaxial stresses to an equivalent uniaxial stress using the Von-Mises method [START_REF] Mises | Mechanik der festen Körper im plastisch deformablen Zustand[END_REF].

First, using the elasticity equations for a unit length cylinder, the radial, circumferential and axial thermal stresses are calculated, respectively [START_REF] Hetnarski | Thermal stresses -Advanced theory and applications[END_REF]:

( ) = ⋅ 1 - [ - 2r - 1 ( ) ⋅ ] ( ) = ⋅ 1 - [ + 2r + 1 ( ) ⋅ -( )] ( ) = ⋅ 1 - [ -( )]
The average temperature is given by:

= 2 - ( , ) ⋅ ⋅
The integration is made in the radial direction. For example, a temperature difference of about 100 °C between the surface and the average temperature, typical for warm start up, causes a thermal stress of about 360 MPa, without taking into account the concentration stress factor.

The stresses due to the rotation of the rotor spinning at 3000 rev/min (50 Hz) for the HP and IP stage are shown in Table 1 and are obtained with the following equations [START_REF] Vullo | Rotors: stress analysis and design[END_REF]:

= [1 -( ) ] ( 3
+ ) 8 = [1 - 1 + 3 3 + ( ) ] ( 3
+ ) 8 Table 1. Centrifugal Rotor Stresses (MPa) HP (d = 725 mm) Surface ( ) ≃ 18
HP (d = 725 mm) Center (0) ≃ 42 (0) ≃ 42
IP (d = 810 mm) Surface ( ) ≃ 22
IP (d = 810 mm) Center (0) ≃ 52 (0) ≃ 52
The blades create also a radial stress at the surface which is taken into account using [START_REF] Vullo | Rotors: stress analysis and design[END_REF]:

( ) = = ( + 2 ℎ ) ℎ 2
Where , , N and ℎ are, respectively, the density of the material, the volume, the number and the width of the blades, is the distance of the center of gravity of the blades from the rotation axis; and are, respectively, the radius and thickness of the shroud. The shroud is an outer ring that surrounds each array of blades, having the task of stiffening them against vibrations. It is also the surface on which the seals work for reducing the escape of steam.. The blade's shape was approximated to a parallelepiped. The data taken into account for the blade centrifugal effect are reported in the Table 2.

= - - 1 - = - - 1 +
The maximum pressure acting on the rotor is 126 bar and generates compressive stress at the surface of the rotor (-12 MPa), even if the steam enters the machine the first time with pressure higher than 63 bar resulting in a -6 MPa stress.

It can be easily noted as the thermal stresses are a lot higher than the stresses produced by both pressure and centrifugal forces, weighing for about the 95% of the overall stress.

Stress Concentration Factor

The effect of stress concentration has been taken into account at the blades' points of insertion on the rotor surface by introducing an appropriate theoretical stress concentration factor in elastic range K [START_REF] Pilkey | s Stress Concentration Factors[END_REF] which involves the radius of the groove of the first stage of blades. For the Tirreno Power turbine rotor it was assumed K = 2.8 for both the HP and IP sections. The assessment of the life consumption associated with start-up has been carried out through to the low-cycle fatigue theory. Manson Coffin curves have been used, according to the type and operating conditions of the alloy rotor.

Influence of Plastic Strain on the Stress Concentration Factor

In low-cycle fatigue (LCF), assessments are performed using strains. To calculate the actual value in the notches, where the plastic state is reached, it has been introduced a strain concentration factor K by means of the Neuber's approach [START_REF] Neuber | Theory of Stress Concentration for Shear-Strained Prismatical Bodies with Arbitrary Nonlinear Stress-Strain Law[END_REF]. The linear elastic model for the evaluation of the strain provides a lower result than the real one.

After the calculation of the nominal stress σ , the value of K is obtained by intersecting the Neuber's hyperbola, having equation σε = K K σ ε = K σ ε and the cyclic stress strain curve of the material.

The result of the evaluation has been represented in Figure 4 for the specific material in the red dashed curve, as a function of the normalized nominal strain range. The trend has been compared with the results proposed for similar materials [START_REF] Viswanathan | Damage Mechanisms and Life Assessment of High-Temperature Components[END_REF].

Multiaxial Loading Conditions

Once the type of fatigue cycle is evaluated, the mean and alternate components have been identified for the three kind of stresses acting on the rotor. These have been combined into equivalent stresses, according to the von Mises yield criterion, obtaining the values σ and σ . This criterion can be applied since the rotor is subjected to centrifugal and thermal loads that act always along the principal directions.

Influence of the Mean Stress

Two approximate methods have been used to take into account the presence of the mean component of the fatigue cycle. They have been used to define an equivalent stress which allows using the Manson Coffin's curves.

The first one exploits the nominal results obtained in the elastic range, amplified by the concentration coefficient of deformation earlier discussed. The equivalent stress is simply calculated as half the difference between the maximum and the minimum value of the fatiguing cycle.

The second method is taken from high cycle fatigue where the mean stress has a linear influence on the fatigue life. The following equation is used, where s is a safety factor (value = 1.2) introduced to consider the uncertainty on the strength characteristics of the material, obtained by [18], since the thermal treatment of the rotor is not known.

= 1 -

In both cases, the amplitude of strain is obtained as

∆ = 2 •
Since the approximate estimates are affected by uncertainty, between the two obtained results it has been always used the more conservative one.

Low Cycle Fatigue Curves

For appraising the maximum number of allowable cycles, related to a given deformation amplitude , the Manson Coffin's curves in the Figure 5 have been used [START_REF] Viswanathan | Damage Mechanisms and Life Assessment of High-Temperature Components[END_REF]. The curves also consider the effects of hold time, i.e. the possible residual plastic strain remaining after intense maneuvers.

For the sake of safety, only the curves referred to the maximum temperature reached during the maneuvers have been considered. Such approximation is conservative, since the maximum strain is not necessarily reached at the maximum temperature.

In rotors subjected to thermal stress, hold times last about 3 hours (i.e. the time of the thermal transient); various authors [START_REF]task: Thermal and Stress Analysis of Rotor[END_REF] [18][19] [START_REF] Timo | The operation of large Steam Turbines to limit Cyclic Thermal Cracking[END_REF], however, suggest to use the curves with 24 hours "hold time" only for the high strain region (number of cycles less than 500). This choice has been applied for the entire operating range, acting in a conservative manner.

CYCLIC LIFE EXPENDITURE CURVES CALCULATION

Description

The strain intensity due to thermal stress depends on the magnitude and gradient of the temperature variation, the heat exchange coefficient, the mass of the rotor and its thermal properties, as well as the presence of notches that can concentrate stress/strain.

The CLE curves give an indication on the consumption of life associated to a given operation.

They also indicate the region to be avoided during start-ups, or the conditions which not to operate in.

CLE curves are parametric ones obtained by connecting points and are function of:

 total range of temperature and its gradient  stress concentration factor  rotor diameter For a given value of the diameter and the stress concentration factor, such curves allows to:

 knowing in advance the reduction of turbine's life connected to a given maneuver or gradient of temperature change  making choices on the reduction of warm up and load times, and then on the increase of temperature gradients that the rotor will have to endure  choosing whether to sacrifice some of the useful life of the rotor to gain greater operating flexibility

Realization

Depending on the type of start-up or performed maneuver, a temperature gradient (linear loading ramp) that simulates the one detected by the starting probes is imposed to the RSE. The algorithm ensures then to calculate the temperature on the axis and the mean temperature, thus allowing the evaluation of the stress components due to temperature, pressure and rotation.

The equivalent stress is then calculated through the Von Mises criterion, using an appropriate value of K (which is greater than K for stresses inside the plastic range while is equal to K in the elastic range, as said above).

Amplitude of strain is then calculated with the methods previously cited, in order to use Manson-Coffin's curves for low cycle fatigue.

Once the number N of life cycles is obtained, its inverse value is multiplied by 100 thus defining the specific percentage life consumption for each maneuver. It can be represented by a point in a graph having the temperature change on the abscissa, the temperature gradient on the ordinate, fixing the stress concentration factor and the rotor's diameter.

Repeating the process, changing each time the total temperature variation and the slope of the loading ramp, different values of N associated with the maneuvers are obtained. They are reported in our graph to obtain a set of iso-consumption life curves. The CLE of the High pressure turbine sections (HP) are shown in Figure 9.

For example, it is possible to observe that life consumption equal to 0.01% may be caused by a temperature change of 100 °C performed at 250 °C/h. In other words, the rotor might tolerate up to 10000 cycles before cracks nucleation.

While increasing the temperature difference of the rotor, to keep constant the life consumption it is necessary to reduce the temperature gradient.

If this action is not made, the curve moves upward with a consequent increase of percentage life consumption.

These CLE curves can be looked up to predict the effects that a certain type of maneuver may cause to the rotor, allowing to choose in advance the best way to operate the unit start up.

To show how a number of cycles N has been obtained and then the identification of a point in the graphic ∆T -∆T/∆t, a warm start up operation performed in Tirreno Power plant is shown (Table 3). It's characterized by an initial rotor temperature of 345 °C, reached 40 hours after the turbine stop (for example, during a weekend).

The data shown in the graphs below (Figures 6, 7 The first graph (Figure 6) shows the trend of the powers generated by the gas turbine and by the steam turbine. It shows also the trend of the temperatures calculated by RSE.

In the second graph (Figure 7), the nominal stresses on the outer surface of the rotor have been calculated using the discussed approximate methods. They are deprived of the contribution of the stress concentration factor to highlight the stresses due to the thermal transient.

The stress components are expressed in the three principal directions, to show their mutual intensity and conserve their sign, in order to understand whether the stress is tensile or compressive.

Finally, in the third graph (Figure 8), it's possible to see the assessment of the equivalent stresses through the use of the Von Mises yield criterion, on the outer surface and on the axis of the rotor.

Observing the shut down it's possible to note that the beginning of the cooling occurs slowly, thus generating thermal stress not exceeding 40 MPa.

This allows stating that even considering as fatiguing cycle the stresses caused by start up and shut down, a fully-reversed loading condition is not got. Therefore it is possible to consider a zero-maxzero profile loading condition. Evaluating of the maximum stresses during start up and applying the approximate method, the calculation of the fatigue life has been performed for this maneuver. The impact of combined cycle startup on the steam turbine life was assessed on the basis of the actual operational data and a simple but rigorous approach, evaluating the Cycle Life Expenditure (CLE) Curves. The data were used to estimate the yearly average life expenditure. The algorithm was then implemented on the Power Plant Distributed Control System (DCS).

The construction of the CLE curves was the goal of the work allowing to predict the consumption of life associated with different operation modes.

Further development will be the exploitation of the designed methodology for optimizing the maneuvers that generate thermal transients.

Knowing in advance the damage caused to the rotor of the turbine, allows to manage the speed of start-ups with greater awareness. As mentioned, this grants the flexibility required by today's energy market: the goal is to shorten the time required for the operations, even at the expense of a higher -but expected-damage with the benefit of improving the competitiveness and efficiency in service delivery.

However, the efficiency might be improved reducing the power consumption and the time required for starting-up, for equal spent life. The better understanding of the effects of the maneuvers will also reduce the risk of damage or failure that may suddenly stop production, planning the maintenance in relation to the calculated cumulative damage, as well.

The work done can be improved with the use of a finite element numerical model for both the calculation of the thermal loads distribution and of the stresses in the rotor. This would make it possible to take into account the specific geometry of the insertion region of the blades, giving more accurate results on the local stresses and deformation in those points. In addition, a more detailed characterization of the fatigue life to the particular study conditions will be the decisive step to achieve trustworthy final results.

ACRONYMS

CCPP

Figure 2 .Figure 3 .

 23 Figure 2. Temperature Comparison for a step change in the surface temperature

Figure 4 .

 4 Figure 4. Strain Concentration Factor

Figure 5 .

 5 Figure 5. Low Cycle Fatigue Curves for Cr-Mo-V Rotor Steels at Approximately 540°C [5]

, 8 )

 8 have been directly recorded from the combined cycle's control unit and concern only the HP turbine.

Figure 6 .Figure 7 .Figure 8 . 3 .

 6783 Figure 6. Temperatures and Powers

Table 2 .

 2 Blades Mass and Geometry Data

				Shroud Radius [ ]		0.402	0.511
		HP	IP	Shroud Thickness [	]	3	3
	Density [ / ]	7800	7800
	Distance between Blade's	0.384	0.458
	Centroid and Rotor Axis [ ]
	Blade Mass [ ]	0.131	0.305
	Blade Number	103	103
	Blade Width [ ]	0.03	0.03

  Combined Cycle Power Plant CLE

		Cyclic Life Expenditure
	DCS	Distributed Control System
	EOH	Equivalent Operating Hours
	GT/ST Gas Turbine / Steam Turbine
	HP/IP	High Pressure / Intermediate Pressure
	OEM	Original Equipment Manufacturer
	OH	Operating Hours
	RSE	Rotor Stress Evaluator

Start Up

Yearly Life Consumption Estimation

Moreover, experimental data for several startups were used to evaluate the actual life expenditure over one operating year of the Combined Cycle. The generalized information were combined through the Miner-Robinson rule [START_REF] Viswanathan | Damage Mechanisms and Life Assessment of High-Temperature Components[END_REF] taking into account also the creep effect due to high temperature normal operation (creep-rupture life under the nominal operating conditions is 300,000 h [START_REF] Viswanathan | Damage Mechanisms and Life Assessment of High-Temperature Components[END_REF]). The Table 4 shows the different start ups' contributions to the life expenditure of the rotor in one year of average cycling operations: as effect of the low cycle fatigue a yearly consumption of 5.05% was calculated (as sum of the terms in the last column). The creep effect, considering an average of 3000 OH per year, adds 1% of yearly life expenditure (3000/300,000). The two cited contributions are simply summed up (6.05%) and the inverse result represents the number of years until defects nucleation (17 years).

CONCLUSION

The activity led to quantify the effects of start-ups and of the load variations on the steam turbine rotor


Ayhan Nazmi İlikan

Erkan Ayder

Effect of the Computational Domain Selection on the Calculation of Axial Fan Performance

Keywords:

In this study, the aerodynamic performance of a jet fan (considered as a free-inlet free-outlet axial fan which does not contain any duct at the upstream or downstream) is obtained by using CFD (Computational Fluid Dynamics) method. The numerical calculations are performed by using the commercial software ANSYS CFX. Three different inlet and outlet computational domain shapes that are widely used for simulations of axial flow fans are evaluated. In the first method, the fan is modeled inside a circular pipe. The second approach does not contain the domain of a pipe at the inlet and the outlet but large inlet and outlet domains connected instead directly to the fan casing. The third method is similar to the second approach except a bellmouth located at the inlet of the fan that is not present in the second approach. The aerodynamic performance of the fan at atmospheric conditions is obtained by three approaches and the results are compared to each other and to the ones obtained experimentally. The results show that the simulation of a jetfan as if it works in a pipe brings an error of 45 % in flowrate which can be decreased to 25 % by placing a bellmouth at the intake of the fan casing.

INTRODUCTION

Computational fluid dynamics has become very popular in last 20 years since the performances of personal computers are being increased enormously since that time. Thanks to the improvements of turbulence models and solution algorithms, turbulent, 3D and unsteady flows inside turbomachines are being solved more correctly day by day. For axial turbomachines, one of the most popular methods is placing the rotor and/or the stator in a fluid domain which simulates the fluid inside a pipe and applying proper boundary conditions [START_REF] Huang | An optimal design for axial-flow fan blade: theoretical and experimental studies[END_REF][START_REF] Shankaran | Validation of an Advanced Fan Model With Multiple Reference Frame Approach[END_REF][START_REF] Sahili | 3-D Modeling of Axial Fans[END_REF][START_REF] Fogal | Theoretical and experimental study of agricultural spraying using CFD[END_REF][START_REF] Dwivedi | CFD Analysis of Axial Flow Fans with Skewed Blades[END_REF][START_REF] Zhu | Experimental and Unsteady Numerical Investigation of the Tip Clearance Noise of an Axial Fan[END_REF]. In this method, the application of uniform inlet velocity boundary condition is popular even if in some studies inlet velocity profiles are applied [START_REF] Rábai | Validation of a computational fluid dynamics method to be applied to linear cascades of twisted-swept blades[END_REF]. However, in most cases, the shape of this profile is unknown, so uniform inlet velocity boundary condition is more popular than the latter one. On the other hand, not all turbomachines work inside a duct. Some axial turbomachines such as house ventilators or jetfans sucks air from open atmosphere and ejects the flow again directly to the large atmospheric area. In these cases, the inlet and outlet conditions would affect the incoming and outgoing air and these conditions would be different than those of a fan working inside a duct. This condition is also true for jetfans working inside short ducts. Therefore, some CFD based studies take large domains at the inlet and the outlet of the fan into account and include it in the computational domains [START_REF] Guedel | Energy Efficiency of an Axial Fan for Various Casing Configurations[END_REF][START_REF] Akturk | A computational and experimental analysis of a ducted fan used in VTOL UAV systems Turbomachinery Conference[END_REF]. The standard testing of the free inlet-free outlet fans are performed mostly according to AMCA210 standard [START_REF]Laboratory Methods of Testing Fans for Certified Aerodynamic Performance Rating. Air Movement and Control Association[END_REF]. In that standard, the fan is tested in conditions in which the exit of the fan is open to a large chamber and the fan sucks directly from the atmosphere. The manufacturers obtain real flowrate by these tests but generally, during the design phase, they perform CFD simulations by placing the fan inside a pipe which would bring some error in the calculation of the flowrate through the fan. The aim of this study is to investigate the effect of the computational domain shape on the flowfield and global parameters (flowrate, total pressure rise, efficiency, etc). To do so, the solid model of a jetfan and the flowrate through that fan that is obtained in an AMCA210 test chamber is provided from the manufacturer. The CFD calculations are performed in three different computational domain shapes. In the first case, the fan is placed inside a pipe; in the second case, large domains covering the fan geometry is added to the computational domain and the third geometry is similar to the second one except an inlet bellmouth that is used to provide uniform inlet velocity conditions.

Numerical Model and Procedure

The fan used in this study is a reversible axial jet fan used in smoke exhaust systems. The manufacturer provides the opportunity to disassemble the long casing (a small circular duct) of the fan from the short casing which covers the rotor and the motor. In this study, only the domain which covers the rotor and the electrical motor (named "core part" in this study) is considered. The specifications and the solid model view of the fan are shown in Table 1 and Figure 1, respectively.

Computational Domain and Grid Generation

The first computational domain (namely 1 st case in this study) consists of the core part of the fan and the axial extensions that represent pipes at the upstream and the downstream of the fan. The axial lengths of these pipes are chosen as 1D where D refers to the tip diameter of the fan (Figure 2). The second computational domain (namely 2 nd case in this study) consists of large domains that cover the core part of the fan. These domains are thought to represent the physics of the flow at the opening regions of the fan to the atmosphere, more realistically. The axial lengths and the diameters of these large domains are chosen this time as 2.5D and 4D, respectively (Figure 3).

Figure3.The computational grid composed of the large domains (2 nd and 3 rd cases)

The third approach (namely 3 rd case in this study) is similar to the second case. The only difference is the presence of a bellmouth shaped geometry at the inlet of the casing of the fan. The bell mouth region in the third approach and the sharp entrance region in the second approach are shown in a closer view in Figure 4. In all the calculations, only one rotor blade and one motor support plate passages are modeled since the flow is axisymmetric in the direction of the rotation. These support plates are kept in the numerical study for comparison with experimental results and should not be confused with stator blades. Since they are not designed as stator blades which follow the absolute flow at the exit of the rotor blades and orient it to the axial direction in order to convert excess kinetic energy due to swirl in to the pressure head. As the number of the rotor and support plates are different (6 rotor blades and 4 motor support plates), the extents of the inlet and rotor domains in theta direction cover sectors of 60°, while the motor support plates and the outlet domains, 90°. The difference of the thickness of the sectors is carried out by the "stage" interface that will be explained in the further sections. The computational grid is generated by ANSYS Mesh. Hexahedral and tetrahedral elements are both used where applicable. In the core part of the fan, tetrahedral elements are chosen due to the complexity of the geometry. The generated mesh for this part of the model is kept the same in all the three computational domain configurations. The rest of the domains (inlet and outlet pipes for the first case and large domains for the second and third cases) are meshed by hexahedral elements to decrease the computational time. Inflation layers are added to all solid surfaces (hub, shroud and the blades of the rotor and the electrical motor; pipe walls) and kept the maximum y + values of these surfaces around 1 to capture the flow physics inside the boundary layers. A section that shows the mesh around the blade and a list of the number of elements related to different domains are shown in Figure 5 and Table 2, respectively. Figure5.The mesh around the rotor blade

The Numerical Algorithm

The flow is modeled as 3D, incompressible, steady and fully turbulent. The finite volume solver ANSYS CFX [START_REF] Ansys | ANSYS CFX-Solver Theory Guide Release 14[END_REF] is used for the calculations of the flowfield. A second order accurate scheme is chosen for the convection equations. k-ω SST turbulent model [START_REF] Menter | Two-Equation Eddy-Viscosity Turbulence Models for Engineering Application[END_REF] is used for the closure problem during RANS calculations since it is known as powerful to predict the flow separation [START_REF] Bardina | Turbulence Modeling, Validation, Testing and Development[END_REF]. y + values are kept around 1 near the solid walls to benefit from the near wall treatment for low-Re number computation of the turbulence model. Air at 25°C is used in all the calculations as the working fluid.

Boundary Conditions

The frame change at the interface between the rotational domain and the stationary domain is chosen as the "stage" model which performs a circumferential averaging of the fluxes through bands on the interface. Since the angle of the section of the rotor and the support are not the same (60° and 90°, respectively), an intersection algorithm provided by the "stage" option provided by the code, called "specified pitch angles" is used at this interface. This option provides to specify the pitch angles on two sides of the domain interface. The pitch change adjustment is performed by applying average values obtained from the upstream side of the interface to the downstream side. The rotational speed of the fan is imposed to the rotational domain and the rest of the domains are kept stationary. The solid surfaces which belong to the stationary and rotational domains are modeled as stationary and rotating walls, respectively.

Counter rotating wall boundary condition with the same magnitude but the opposite direction to the rotation of the blades is imposed to the shroud surface of the rotating domain to make this surface fixed in stationary frame. In the first case, the "uniform inlet total pressure" boundary condition is imposed to the inlet of the pipe. The total pressure value in this surface is set as zero since the fan in consideration works in atmospheric conditions. The "opening pressure" boundary condition is imposed to the outlet boundary of the pipe of this case and also to the external surfaces of the outlet large domains of the second and third cases. The reason of the usage of the opening type boundary condition instead of outlet static pressure option is about the direction of flow. Static pressure option in CFX does not allow reverse flow at outlet boundaries. If the code calculates reverse flow at these boundaries, it imposes artificial walls which are not physically true. This problem is solved by opening pressure boundary condition which allows bidirectional flow. During the calculations, this opening pressure is kept at atmospheric conditions (0 Pa gage pressure). The conditions that define turbulence properties at the "inlet" and "outlet" surfaces are set as "zero gradient" that is recommended by the code when opening boundary condition is used in one of the boundaries. Since the flow is modeled as periodic, periodic boundary conditions are imposed to side surfaces.

RESULTS AND DISCUSSION

In this section, the results obtained from the calculations will be presented by means of charts, contours and tables obtained at the inlet and outlet of the fan, as well as the passage domain. The flow coefficient of the 2 nd case has an error of 1.5 % only since the physical modeling of this case is the closest one to the experimented working conditions of the fan. On the other hand, the unrealistic uniform entry conditions at the upstream of the 1 st case cause the prediction of the flowrate to be almost 45 % higher than the one obtained by the experiment. The third case shows that the bellmouth shaped geometry at the inlet provides to increase the flowrate by 16 % compared to the 2 nd case. This ratio is in accordance with Bleier's claim [START_REF] Bleier | Fan handbook: Selection, application, and design[END_REF] that is the possibility of increase of flowrate up to 15 % owing to the inlet bellmouth. The second important conclusion is related to the magnitude of the error in case of a simulation of a fan with bellmouth by using the first approach (modeling the jetfan with a bellmouth in a pipe).

A comparison of the flow coefficients of 1 st and 3 rd cases in Table 3 shows that in such a case, that type of a simulation would over predict the flowrate by 25 %. The total pressure rise coefficient shown on the Table 3 is defined as the difference between the mass flow averaged total pressure at the exit plane of the fan casing (shown in Figure 6) and the stagnation pressure at atmospheric conditions divided by the dynamic pressure based on the peripheral velocity at the tip section of the rotor blade. The first and the second cases show total pressure rise results close to each other despite the remarkable difference between their flowrates. The reason is related to the losses inside the jetfan. The bellmouth provides nearly uniform conditions at the inlet of the fan that decrease intake losses. However, the intake losses still exist unlike the 1 st case. On the other hand, low intake losses result in high mass flow which in turn causes high swirl at downstream of the rotor that causes again higher losses. The balance of these opposite effects and the mixing at the exit domain take role and finally, the 1 st case show a higher flowrate compared to the 3 rd one despite the same total pressure rise. This means that the intake conditions cause the characteristics of the fan to be changed. In the second case, the flow at the upstream of the rotor blade is blocked because of the poor intake conditions. This event causes a remarkable change in the characteristics of the fan.

In this case, because of the related severe pressure losses, the total pressure rise coefficient and the flow coefficient have both low values. The thrust is directly related to the volume flowrate that is why the closer thrust value to the experimental one is shown again by the 2 nd case.

Inlet

Figure 7 and 8 show the circumferentially averaged flow coefficient and the total pressure rise coefficient distributions in radial direction, respectively, obtained at the rotor inlet plane (Figure 6). In the 2 nd case, the magnitude of the axial velocity is decreased starting from the 70 % of the nondimensional radius because of the vena contracta that occurs due to the poor intake conditions.

Figure6.The planes used for the calculations of the parameters

Consequently, backflow occurs between 90 % and 100 % of R/Rtip. The reason of the higher axial velocity in the 2 nd case up to 70 % of R/Rtip is about the strong vortex that occurs at the inlet of the blade that will be shown in subsequent sections of the paper. The reason of the higher total pressure at the tip region in this case that is shown in Figure 8 is related to this strong backflow originated from the exit domain of the blade. On the other hand, the bellmouth of the 3 rd case seems to remove that blocked region by providing more uniform flow conditions at the inlet of the rotor. The magnitude of the axial velocity is lower but the total pressure is almost the same in all along the blade compared to the 1 st case as already discussed above.

Passage

Figure 9 shows total pressure contours and the absolute velocity vectors projected on a meridional plane. A perspective view of this plane can be found in Figure 6.

The results show the effectiveness of the bellmouth. Almost similar inlet conditions occur in the 1 st and 3 rd cases. The bellmouth provide the low total pressure region at the inlet to remain narrow and close to the casing whereas in the second case, a strong vortex occurs at the inlet of the blade. The flow conditions at the position of the fan casing outlet plane (shown in Figure 6) are also remarkable. At the position of that plane, unlike the results up to now, the 2 nd and the 3 rd cases show similar contours in Figure 9. This time the results of the 1 st case is quite different since the casing of this case is elongated in axial direction that prevent the mixing of the flow leaving the jetfan with the atmosphere.

In Figure 10, the pressure contours combined with the relative velocity vectors projected on a blade-to-blade surface are shown. This surface is close to the tip region of the rotor blade corresponding to R/Rtip=0.95 (93 % of the span starting from the blade root). Most of the relative velocity vectors in 2 nd case seem to be directed to inlet of the fan that means backflow. Because of this backflowswirl combination, the fluid does not flow around the pressure and suction surfaces of the blade but it hits the pressure surface somewhere far away from the leading edge. This stagnation point divides the flow into two regions. The one which is directed to the exit plane flows around the pressure surface while the other one is separated near the leading edge. This results in the loading of the blade to be decreased in this section of the blade (Figure 11b). The stagnation points of the 1 st and 3 rd cases do not occur on the leading edge but are on the pressure side at a position closer to the leading edge than the one of the 2 nd case. The contours and velocity vectors are similar in 1 st and 3 rd cases which have both massive separations in the suction side. The attached region on the suction side of the 1 st case seems to be larger than the one of the 3 rd case which can be also seen in the Cp distribution chart in Figure 11b. It is well known that the area enclosed by the pressure curves of the suction and pressure sides give the magnitude of the work done by the blade. Thus, from the charts given in Figure 11b, one can conclude that the work done by the blade is much more in 1 st and 3 rd cases compared to the one of the 2 nd case. On the other hand, Figure 11a shows the pressure distribution on the suction and pressure sides of the blade at R/Rtip=0.7 corresponding to 57 % of the blade span starting from the blade root. In this region, the loading of the three cases are more similar compared to the ones of the R/Rtip=0.95. However, one can notice that the suction side of the 2 nd case is less loaded which is again a consequence of the vortex at the inlet region and the related backflow at the upper part of the blade that affects the upper-mid span of the blade as well.

Outlet

In Figure 12 and 14, circumferentially averaged flow coefficient and total pressure coefficient in radial direction are shown. The charts on Figure 12 and 14 are obtained behind the rotor and behind the exit of the fan casing, respectively. The results show that the poor inlet conditions affect the loading of the blade which in turn cause the total pressure rise and the axial velocity to be decreased in all along the span.

Figure 13 shows the axial velocity contours just behind the rotor. A separated region is found in the suction surfaces near hub region of all the three cases. The axial velocity contours are similar in that region (can be seen also in Figure 6) which means that the inlet conditions studied in this paper do not affect the flow regime near hub region. However, at the upper radii, low axial velocity values can obviously be seen in the 2 nd case that affects the flow in midspan region as well. The 3 rd case shows that the axial velocity distribution in upper radii is improved so that low velocity region is confined to the region near casing. 14a and 14b shows that in all the three cases, nearly uniform axial velocity distribution in radial direction is provided at the exit plane of the jetfan. Negative total pressure at the wake region of the fan motor of the 1 st case in Figure14b differs from that of the 2 nd and 3 rd cases. This is a consequence of the mixing of the fluid with large domains in 2 nd and 3 rd cases whereas there is no such a mixing in the 1 st case.

CONCLUSION

The effect of the computational domain shape on the aerodynamic performance of a jetfan is investigated. Three popular domain configurations are studied and the global parameters as well as the flowfields of these configurations are investigated and the differences of the results are discussed. The results show that simulating a jetfan as if it works in a pipe would bring a considerable amount of error because of the modified inlet conditions as well as the lack of the mixing with the atmospheric air at the exit of the fan. On the other hand, if the jetfan possess a bellmouth at the inlet to provide uniform inlet conditions, the error when a simulation is performed as if the fan works in a pipe is decreased considerably from 45 % to 25 %.

  Figure1. Jet Fan

  computational grid composed of the fan and the pipes by (a) internal (b) external view (1 st case)

  comparisons of the entrance regions of the second (a) and third (b) approaches

Figure7.

  Figure7.Flow coefficient distributions in radial direction at the rotor inlet plane

Figure9.

  Figure9.Total pressure contours and absolute velocity vectors projected on a meridional plane: (a) 1 st case, (b) 2nd case, (c) 3 rd case.

Figure10.

  Figure10.Pressure contours and relative velocity vectors on a blade-to-blade plane: (a) 1 st case, (b) 2 nd case, (c) 3 rd case.

Figure13.

  Figure13.Axial velocity contours at the rotor outlet plane: (a) 1 st case (b) 2 nd case (c) 3 rd case

Figure

  Figure14aand 14b shows that in all the three cases, nearly uniform axial velocity distribution in radial direction is provided at the exit plane of the jetfan. Negative total pressure at the wake region of the fan motor of the 1 st case in Figure14b differs from that of the 2 nd and 3 rd cases. This is a consequence of the mixing of the fluid with large domains in 2 nd and 3 rd cases whereas there is no such a mixing in the 1 st case.

Figure14.

  Figure14. (a) Flow coefficient and (b) total pressure coefficient distributions in radial direction behind the fan casing outlet plane.

Table 2 .Mesh information Element type # of elements

 2

	Rotor	Tetrahedral	2469224
	Electrical Motor Support	Tetrahedral	2750409
	Plates
		Hexahedral	2199370
	1 st case (total)	Both	6175043
	2 nd case (total)	Both	7241773
	3 rd case (total)	Both	7419003

Inlet & Outlet pipes (1 st case) Hexahedral 955410 Inlet & Outlet large domains (2 nd case) Hexahedral 2022140 Inlet & Outlet large domains (3 rd case)

Table 3

 3 shows flow coefficient Φ, total pressure rise coefficient Ψ, thrust T, total efficiency ηt values obtained from the simulations realized under atmospheric conditions. The flow coefficient and thrust values are compared to the experimental ones that are already shown in Table1, obtained by the manufacturer in an AMCA210 test chamber.

Table 3 .

 3 Performance Parameters

	Φ	ΔΦ (%)	Ψ T [N] ΔT(%) ηt (%)
	1 st case 0.381 +43.2 0.229 28.2 +101.4 63.2
	2 nd case 0.262 -1.5 0.169 13.3 -5.0 39.8
	3 rd case 0.305 +14.7 0.226 18.2 +30.0 52.0

ACKNOWLEDGMENTS

The authors would like to thank to Bahçıvan Elektrik Motor San. ve Tic. Ltd. Şti. for providing the geometry and the experimental data of the axial flow fan.


Tobias W Zimmermann
email: zimmermann@ikdg.rwth-aachen.de

Oliver Curkovic

Manfred Wirsum

Andrew Fowler

S Y Mpo

O Tatin

Influence of adjusting control accuracy on pressure probe measurements in turbo machines

Keywords:

This paper presents a new design of a probe adjusting device intended to position pressure and temperature probes in a flow field. 5-hole , 3-hole and temperature probes can be moved in radial direction and freely rotated about their axis. The high actuation accuracy of 3.9µm in radial direction and 0.09°in angular position is validated in a 2-stage-turbine test rig which is installed at the Institute of Power Plant Technology, Steam and Gas Turbines, RWTH Aachen University.To meet the challenge to calculate the efficiency of a turbo machine which is mainly influenced by the temperature, all probe adjusting devices are positioned simultaneously and controlled by the measuring acquisition system (MAS) so that the same radial position in each stage is measured at the same time. For this purpose a new program has been developed to synchronize actuation and measurement. The slim design of 60mm width allows measurement between the stages of turbomachines with small axial distances between vane and blade. In addition a CFD/FEA shows how the design and combination of materials compensate the thermal expansion of the engine during operation. This allows a minimal safety distance of 0.2mm between rotor and probe to enable measurement as close to the physical boundary as possible. The actuation accuracy is demonstrated with pressure, temperature and angle distribution plots. It is also shown that the resolution of the measuring points, and therefore the actuation distances, has a large impact on the flow field analysis and should be set as high as possible. However the measuring time has to be taken into account

INTRODUCTION

Taking the increasing efficiency of steam turbines into consideration, the process parameters need to be enhanced. This thermodynamic enhancement results in a higher density respectively a lower volumetric flow at the inlet of a turbine. For this reason the airfoils, which still need to be developed are indicated by small blade height and low height/chord-ratios. With the usage of shrouded blades the leakage flow losses are reduced [START_REF] Schwab | Entwicklung einer Instrumentierung für Dampfturbinen zur dreidimensionalen Bestimmung von Strömungsfeldern und Analyse von Strömungsverlusten in Leit-und Laufbeschaufelungen mit Deckbändern, Abschlussbericht zum Forschungsvorhaben 0327723E[END_REF]. The flow conditions in such flow channels are very complex, thus a CFD simulation without experimental validation is not suitable to optimize the aerodynamic blade design. For this purpose 5-hole-probes (5HP), 3-hole-probes (3HP) as well as temperature-probes (TP) have to be inserted in the flow channel to determine the radial flow profile across the span. E.g. a 5HP uses five holes arranged as a cross to measure flow angle, Mach number or static pressure calculated by means of pressure differences that can be approximated with calibration matrices gained after calibration in a wind tunnel [START_REF] Bohn | Mehrparametrige Approximation der Eichräume und Eichflächen von Unterschall-und Überschall-5-Loch-Sonden[END_REF], [START_REF] Fiedler | Strömungs-und Durchflussmeßtechnik[END_REF]. All types of probes are positioned with a probe adjusting device (PAD). Modern airfoil designs with tangential end wall contouring (TEWC) impact the flow quite close to hub or tip. To determine any changes within the boundary layer the probe has to be positioned as close as possible to the wall to capture all phenomena. In addition to a precise linear movement that adjusts the radial position within a flow channel, the probe design requires rotation of the face perpendicular to the flow due to measurement deviations at the outer boundaries of the calibration area [START_REF] Nitsche | Strömungsmesstechnik[END_REF]. Therefore the PAD has to ensure a reproducible linear as well as a rotational movement of the probe however the space requirements have to be kept. The determination of the engine's efficiency is mainly influenced by the temperature difference between inlet and outlet. Smallest deviations of the temperature would cause errors of the calculated efficiency. To ensure that all temperature measurements are simultaneously related to the inlet temperature, it has to be ensured that all probes measure at the same position within the flow channel at the same time. The present paper therefore shows the results of the measurements at the inter stage of a 2-stage turbine because the most complex flow structure is expected at this location due to rotor -stator interaction and the re-entering of leakage flows. It is shown how the number of measurements across the flow channel influences the measurement quality. A coarse grid with 13 radial measurements and a fine grid with 56 measurement points are shown in comparison to a CFD prediction. Furthermore the radial adjusting accuracy is evaluated by measuring the same conditions with 3HP and 5HP. The probe heads are different in size respective blockage effect and are calibrated seperately with their own calibration coefficients, however the resulting flow distribution has to be similar. To analyse if the rotational accuracy of the PAD is given, the yaw angle α is discussed, because it is calculated considering the angle derived by the pressure measurement as well as the physical angle of the probe that is turned at each measurement position towards the flow. In addition 3HP and 5HP are rotated differently to avoid systematic errors.

NOMENCLATURE

Gr

DESIGN OF THE PROBE ADJUSTING DEVICE

To meet the introduced challenges of measurements in complex flow fields, a new PAD has been developed. The introduced PAD design allows to move the probe 80mm in radial direction with an accuracy of 3.9µm. Simultaneously, the probe can be rotated around its axis while a precision of 0.09°is realized. To ensure the correct positioning, both parameters are surveyed Figure 1. Assembly of probe adjusting device digitally at the very probe. If the probe is misaligned, the actuator moves until the desired position is reached. The assembled PAD is shown in figure 1. The actuator is placed coaxially to the probe to ensure a slim design of 60mm width. The probe is guided by a sledge that moves along a linear guiding which is designed especially for small tolerances. A connector within the sledge is used to access the pressure tubes or the thermocable of the probes. For the present investigations a 3HP with 1mm height respectively a 5HP and TP with a diameter of about 4mm are used. The position monitoring and the rotary encoder are mounted at the sledge, which ensure to measure the position at the very probe. The sealing of the probe is realised within the adapter flange which also serves to fasten the PAD on the casing of the turbo machine. To ensure the high precision, all components have been manufactured with small tolerances

ACTUATOR CONTROL SOFTWARE (ACS)

As written in the introduction, the flow in turbo machines is complex and the probe head should face the flow almost perpendicular before the measurement starts. The required simultaneous movement of several PADs has also to be taken into account thus a new program has been developed in LabView ® .

Figure 2 shows the user interface (GUI) of this new program. Each measurement plane (MP) has its own window which contains two probe monitoring units.The IP of the controller varies for each MP thus the communication protocol is isolated and misguided data packages are outlined. This feature is important to avoid the collision of probe and rotor or casing. Due to the high precision of the PAD, the probes have to be calibrated one time and are then positioned repeatable with a safety distance of about 0.2mm to the hub. A mesh for each probe can be loaded that contains the upper and lower limit of the flow channel and the angle contributed to each radial position within. This ensures a proper measurement of the flow due to a flow facing probe. The current radial position and probe angle is shown in diagrams below the progress bar which displays the status of the current measurement. If necessary, the user may adjust the probe manually. After starting the measurement software for each stage, all probes are moved to the hub. The ACS communicates the status to the MAS that triggers the next probe position after finishing the measurement. This procedure ensures a quick and reproducible monitoring of the flow within the channel.

TEST RIG SPECIFICATIONS

The test rig is a axial 2-stage test turbine, with shrouded constant section blade design. Tangential end wall contouring is applied on hub and shroud side of the airfoils. Labyrinth seals are used to seal the cavities. Further details according to the airfoil and sealing setup as well as for the uncertainty analysis can be found in [START_REF] Schwab | Numerical and experimental investigation of tangential endwall contoured blades in a 2-stage turbine[END_REF] and [START_REF] Schwab | Experimentelle Untersuchung von umfangsunsymmetrischen Dampfturbinenbeschaufelungen und von Temperaturausgleichsphänomenen an einer 2-stufigen Versuchsturbine[END_REF] .

CALCULATION OF THE THERMAL EXPAN-SION

To ensure that the desired radial position of the probe is not influenced by the thermal expansion of the components due to the heating process, the PAD materials have been chosen carefully. The geometry of the turbine is modelled with the CAD software Autodesk Inventor 2014 © . The model consists of rotor (green), flow channel (blue), casing (cyan), PAD (magenta) and the probe itself as shown in Figure 4. The highest temperature can be expected at the inlet of the turbine thus the maximal thermal expansion is located in this area, too. Therefore the FEA model does not consider blades or vanes. The meshing was performed with ANSYS ICEM CFD 15 © . Five nested O-grids cover the contact surfaces of casing, PAD flange and probe (Figure 5). The specific material properties are calculated as shown in Table 2 [START_REF]VDI Wärmeatlas[END_REF], [START_REF] Grote | Dubbel[END_REF], [START_REF] Böckh | Wärmeübertragung[END_REF] and A CFD/FEA simulation has been carried out to validate the amount of the probe displacement. Figure 6 shows the temperature distribution within all components after heating up the engine to its operating temperature. It is obvious that the PAD behaves similar to a fin thus the heat is transferred through it more quickly than through the casing. The safety distance of the probe head to the hub is about 0.2mm which is also a significant reason to investigate the thermal expansion due to a collision in case the probe touches the rotor. Table 3 shows the result of the resulting thermal expansion predicted by the simulation:

NUMERICAL SETUP

To analyse the flow, a CFD model of one blade passage has been generated and simulated with the commercial solver ANSYS ICEM CFD 15 © . Figure 7 shows the flow channel consisting of inlet, outlet and both stages. The simulation area includes labyrinth seals with shrouds and end wall contoured airfoils. The flow channel lies between suction and pressure side of the airfoil to avoid a cut of the end wall contour. The evaluation planes have been set in accordance to the measuring planes in the turbine as described in section 3. Each blade row consists of 55 airfoils. The numerical model is meshed with ANSYS ICEM ©) by means of a structured Hand O-grid. Due to periodicity of a turbine, only one pitch of 6.54 degree is modeled to reduce the computational effort.

A mesh study has been carried out that compares five meshes with 11, 15, 23, 35 and 47 • 10 6 nodes. This study turned out that a mesh with 23 • 10 6 nodes is appropriate to capture all vortices and flow phenomena for the present investigations. The y+ factor is smaller than 2 and SST is used as turbulence model. It switches between the k -and k -ω-model by means of a blending function to take the advantages of each model depending on the calculated area. Its reliability has been proven in several publications [START_REF] Pöhler | Erhöhung des Wirkungsgrades von Turbinenstufen durch Realisierung nicht-rotationssymmetrischer Seitenwandkonturen in den Statorbeschaufelungen[END_REF], [START_REF] Niewöhner | Erhöhung des Wirkungsgrades von Turbinen durch Kombination von nichtrotationssymmetrischen Seitenwandkonturen und dreidimensional gestalteten Beschaufelungen[END_REF] and [START_REF] Germain | Improving Efficiency of a High Work Turbine Using Non-Axisymmetric Endwalls, Part I: Endwall Design and Performance[END_REF]. For each OP, the boundary conditions are derived by measured 5HP data. At the inlet, the total pressure and total temperature distribution is set. The static pressure distribution serves as outlet boundary conditions for the CFD-simulation. Mixing planes are used to switch the rotating and non-rotating systems.

RESULTS AND DISCUSSION

In advance to the redundant tests that are carried out to verify the accuracy of the new PAD, it has to be checked if the resolution of the measurement is high enough to detect all phenomena that occur within the flow. Especially small vortices which are influenced by the TEWC have to be measured in high resolution to learn how they impact the flow. Figure 8 shows the total pressure distribution of the channel in measurement plane MP12 that is located between the two stages of the turbine. This MP has been chosen for all further investigations because the interaction of rotor and stator cause the most complex flow structure effects. Two different resolutions of the PAD grid are compared to the CFD simulation. The coarse grid with 13 radial measurements that was measured in a previous investigation (same experimental setup) with another probe adjusting device is plotted in blue.

The curve in red shows the measurement data with the new PAD and a higher resolution of 56 steps. The span height is the same for both grids. OP1 has been chosen because it represents overload condition and therefore shows more turbulence compared to OP3. Please note that the CFD simulation serves to illustrate the vortex phenomena. However one can see that the CFD prediction shows deviations compared to the test, its quality is sufficient to depict the present flow structure.

The reader may recognise a more detailed distribution of the vortices at 30% and 83% radial height for the fine mesh compared to the coarse mesh. The chosen resolution impacts the flow angle measurement even more than the pressure measurement which is shown in Figure 9. It is obvious that the missing measurement points at 20% radial height do not cover the angle distribution correctly. At this position the flow shows an angle of -40• instead of -35•. Actually a higher deflection is predicted by the CFD which leads to the assumption that the test with low resolution does not detect the flow properly. This result supports the statement that a suitable resolution is as important as the precise positioning to detect all effects within a complex flow field. With this knowledge the resolution for all further tests has been set to 56 steps per radial height.

To validate the positioning accuracy of the new PAD several tests have been performed with the introduced test rig. Two operating points (OP1 and OP3) have been measured two times on several days to ensure that the distribution curves are not influenced by the probe type itself, the measurement has been conducted with a 3HP first and then has been repeated with a 5HP. Two different TP were inserted respectively however the design has been the same.

The measured data shown in Figure 10 -Figure 12 is compared to the CFD results. It has to be taken into account that the data of the 5HP measurement has been used as boundary condition for the CFD simulation, which is plotted in each diagram because the focus lies on the position of flow phaenomena especially close to the hub and tip side where the main impact of the TEWC is expected.

Figure 10 shows the total pressure distribution in MP12. The vortex peaks at 16%, 31% and 83% for OP1 as well as the peaks at 9%, 22% and 88% radial height for OP3 are measured redundantly on both test days. Positions and magnitudes of the peaks are also predicted by the CFD. For OP1 both measurements also show the same distribution. However the CFD matches the experimental data in the center of the flow channel, deviations occur on the casing side. For both OP it has to be mentioned that the CFD predicts peak amplitudes close to the hub which are not measured in the experiment.

The temperature peaks at 20%, 33% and 86% for OP1 and 12%, 30% and 82% radial height for OP3 shown in Figure 11 have been measured reproducibly as well. deviations between both measurement days are below 0.1K. Due to the small changes in efficiency that are investigated nowadays, such amount of deviation has to be ensured to answer questions of efficiency improvements related to new features. Therefore the measurement of all probes in all MP is triggered at the same time and the same span position respectively. The yaw angle has also been taken into consideration because the actuation accuracy around the axis impacts its calculation immediately. It is calculated by summation of the physical probe angle that is set with the PAD and the derived angle which results of the probe pressure measurements. Therefore the probe adjusting accuracy impact the yaw angle directly. One can see that the channel vortex which is located between 10% and 40% radial height causes a massive deflection of the flow in both directions. This deflection requires rotation of the probe towards the flow due to the fact that 3HP and 5HP only work probably up to an angle of approx. ±25 • . Ideally, the probe faces the flow perpendicularly before the measurement is triggered. Therefore a CFD simulation is run before the test and the predicted distribution is used to program the adjusting mesh for each OP. The calculation of the flow angle by pressure differences can be observed online during the test. This measure allows to survey if the adjusting mesh has been set correctly. In case that an angle range of 10 • is exceed, the meshed can be modified to increase the measurement accuracy. The curves in Figure 12 clarify that even changes of about 40• (OP1 20% -30% radial height) can be measured properly and repeatable due to tracking the probe almost perpendicular to the flow at each measuring position. In addition, the radial actuation accuracy has been verified because the head diameter of the 3HP is different to the 5HP design which would result in a radial offset if the PAD owuld not adjust the probes properly. Influence of adjusting control accuracy on pressure probe measurements in turbo machines -9/9

CONCLUSION

Complex flow phenomena in turbo machines require highly resolved measurements within the flow channel to validate numerical simulations and to learn specific details of vortex systems. Probes have to be positioned precisely and redundantly within the annulus to provide this data. Limited installation space of many applications demand compact units to fulfil this purpose. For that reason a new probe adjusting device has been developed at the Institute of Power Plant Technology, Steam and Gas Turbines, RWTH Aachen University that is able to move 3HP, 5HP and TP to a desired position. The small width of 60mm and the high actuation accuracy of 3.9µm in radial direction and 0.09°around the probe's axis characterize this new design.

The material combination of all parts has been chosen with respect not to impact the radial probe position neither to allow a collision of probe and rotor. A FEA/CFD as well as measurements on a 2-stage turbine test rig have demonstrated that an initial distance of 0.2mm probe to hub at ambient conditions changes only by about 0.023mm after heating up the engine to its operating temperature of 363.15K. Even smallest deviations of the inlet temperature could cause errors of the efficiency measurement if other temperature measurements cannot be related to it. For this purpose a new software application has been programmed which allows to move several probes time synchronized to measure e.g. the temperature in each stage simultaneously. This enhances the accuracy of efficiency determination strongly because the temperature drop over stages or the complete engine is measured at the same time.

To ensure that the probe measures accurate, a mesh can be loaded for each probe. It contains the boundary positions on the casing and hub side as well as the contributed angles for each radial position that have been determined via CFD in advance. This procedure allows to rotate the probe almost perpendicular to the flow at each measuring position for a specified OP. Two operating points were measured on different test days with different kind of probes redundantly. The results show that the probe adjusting device actuates the probes as precisely as demanded. CFD simulations have been carried for both OP to allow comparison of the redundant test days. These results show likewise that the position accuracy of the new PAD meets all requirements to survey complex flows in turbo machines 1 1©2016 by General Electric Technology GmbH. All rights reserved. Information contained in this document is indicative only. No representation or warranty is given or should be relied on that it is complete or correct or will apply to any particular project. This will depend on the technical and commercial circumstances. It is provided without liability and is subject to change without notice. Reproduction, use or disclosure to third parties, without express written authority, is strictly prohibited.

Figure 2 .

 2 Figure 2. GUI of actuator control software

Figure 3

 3 illustrates the meridional cross-section of the air turbine schematically. The flow measurement is conducted with a 3HP, 5HP and TP in MP10 (in front of the first stator) and MP12 (inter-stage) and a 5HP and TP in MP22 (behind the second rotor). To measure the flow close to the side walls all probes are positioned with a minimal distance of 0.2mm. The turbine is supplied with air by two radial compressors at an inlet pressure (static pressure) of 3.2 * 10 5 Pa, an inlet temperature (total temperature) of 363.15K and an air mass flow of maximum 14kg/s.The air is pumped in a closed cycle to ensure that changes in environmental conditions do not affect the measurement. The temperature at the compressor and turbine inlet is adjusted by two water driven air coolers. The pressure level in the closed loop is adjusted by a compression load supply and an exhaust valve. The power of the rotor is dissipated using a water brake, mounted in a momentum pendulum with hydrostatic bearings and a maximum performance of 450kW. Via these components the momentum of the rotor can be measured very sensitively by means of a load cell. A honeycomb structure is used at the upstream the first stage to ensure homogenous flow in MP10. The operating points (OP) that are investigated are adjusted by speed while inlet pressure, inlet temperature and pressure ratio are kept constant. OP3 (4775RPM) represents the design point and OP1 (3552RPM) depicts overload conditions.

Figure 3 .

 3 Figure 3. Cross section of the test rig

Figure 4 .

 4 Figure 4. FEA model

Figure 5 .

 5 Figure 5. Meshing of probe head and flow channel

Figure 6 .

 6 Figure 6. Temperature distribution of components

Figure 7 .

 7 Figure 7. CFD flow channel of the 2-stage turbine

Figure 8 .Figure 9 .

 89 Figure 8. Total pressure in measurement plane 12

Figure 10 .

 10 Figure 10. Total pressure distribution in measurement plane 12 for two operating points

Figure 11 . 9 Figure 12 .

 11912 Figure 11. Total temperature distribution in measurement plane 12 for two operating points

Table 1

 1

summarizes the combined standard uncertainties of flow parameters measured with pneumatic multihole probes

Table 1 .

 1 Combined standard uncertainties of flow parameters measured with pneumatic multihole probes[START_REF] Schwab | Experimentelle Untersuchung von umfangsunsymmetrischen Dampfturbinenbeschaufelungen und von Temperaturausgleichsphänomenen an einer 2-stufigen Versuchsturbine[END_REF]

	Flow parameter	Unit	Uncertainty
	α	[°]	0.10339
	γ	[°]	0.15626
	Ma	[-]	0.0008933
	p tot	[mbar]	3.1238
	p st at	[mbar]	3.4108
	T	[K]	0.1299

Table 2 .

 2 Heat transfer coefficients

			Casing Flange	PAD	Probe
	T W	[K]	358.15 323.15 313.15 313.15
	L	[m]	0.487	0.06	0.459	0.099
	Pr	[-]	0.7075 0.7075 0.7075 0.7075
	Gt	[-]	914.24 712.39 19135 1905.6
	Nu	[-]	104	14.69	66.82	23.3
	α	W m 2 K	5.604	6.425	3.821	6.191

Table 3 .

 3 Thermal expansion of the assembly

	Initial Distance	0.2mm
	Rotor	-0.185mm
	Casing	+0.290mm
	Adapter Flange	+0.031mm
	PAD	+0.025mm
	Probe	-0.138mm
	Resulting Distance	0.223mm

ACKNOWLEDGMENTS

The numerical and experimental investigations were conducted as part of the joint research program COORETEC-Turbo in AG Turbo. The work has been supported by the Bundesministerium für Wirtschaft und Technologie (BMWi) (file number 03ET2013) on the basis of a resolution of the German Parliament. The authors gratefully acknowledge AG Turbo and Alstom Power (now General Electric Company) for their support and permission to publish this paper. The responsibility for the content of this publication lies with the authors.