14th Australasian Fluid Mechanics Conference Adelaide University, Adelaide, Australia 10-14 December 2001

Redesigning the Rotor Fan Bthe Cooling of Roxburgh’s H

B. Liddell1, A. Tucker1, I. Huntsman1, M. M

1Department of MechanicalUniversity of Canterbury, Christchurc

2Contact HydroContact Energy Limited, Clyde, 91

Abstract

Since 1995, the level of air flow within the hydro-generators at Roxburgh Power Station has been insufficient to maintain stator winding temperatures below maximum allowable limits from January-April. In the past this has resulted in the de-rating of the machines by 12.5% and substantial loss of revenue. The objective of this research was to implement a cost effective design change to increase the air flow rate past the stator windings without significantly adding to windage losses. A combination of experiment and 2D and 3D computational fluid dynamics (CFD) modelling was used to investigate the existing air flow and highlight possible solution strategies. Based upon the results of these investigations the redesign of the rotor fan blades was chosen as the best solution. The 3D CFD models were then further developed to determine a suitable form for the new fan blades. Calculations made to evaluate the relative performance of the redesigned fan blades indicated a 40% greater pressure boost, a reduction in local viscous losses from 60%-40% of pumping power, but a 10% increase in local windage losses, compared with the original blades. Introduction

Roxburgh Power Station is located on the Clutha River in the Central Otago region of New Zealand and has been owned and operated by Contact Energy since 1996. It has eight 40MW British Thompson-Houston generators that were commissioned in two instalments of four machines in 1956 and 1962. A sectional model of a Roxburgh generator and its 56,000hp Francis turbine is shown in Figure 1. In 1995, it became apparent that the air flow through the generators was unable to maintain stator winding temperatures within their normal 65-75°C operating range from January-April. Since 1995, the over-heating of the stator windings has resulted in the de-rating of the generators from their maximum 40MW power output to 35MW and substantial loss of revenue for Contact Energy. It is also important to limit the maximum operating temperature of the stator windings to control the rate of insulation breakdown and get the expected life from the generators before failure and expensive repairs. Elevated temperatures also increase the electrical resistance in the windings and compound the over-heating problem. The over-heating of the stator windings is the result of a combination of the following factors: • The breakdown of the stator winding insulation which

decreases heat transfer from the conductors. • The accumulation of dust and oil on stator winding and heat

exchanger surfaces reducing convection to the ventilating air. • Sustained high air and river water temperatures in Central

Otago during summer. • An inefficient ventilation system. The gradual heating of the concrete foundations also further compounds the over-heating problem when the generators are run near their maximum output during the summer months.

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as proposed that a simple and inexpensive design change d be made to the generators to increase the air flow past the r windings and solve the over-heating problem. This ification could not cause a significant increase in total age losses above the existing 200kW.

e 1. A sectional model of one of Roxburgh’s eight 40MW British pson-Houston generators with its 56,000hp Francis turbine.

tilation System

generators have a closed-circuit radial-axial ventilation m as described in [1] which has separate radial and axial

ponents. Air flow paths for this ventilation system are shown radial cross-section drawing of the generators in Figure 2. radial ventilation component uses the centrifugal pumping of otor spider and poles to draw air through the rotor. This ides an even flow distribution over the height of the stator. axial ventilation component uses fan blades on the rim of the to drive air between the rotor poles. There is manifold flow een the rotor poles as air is bled off through the stator ducts therefore lower axial velocities and higher pressures near the e of the stator. As a result there is an uneven flow ibution that increases from the ends towards the centre of the r and therefore temperature gradients along the windings. r exiting the stator the air is cooled by eight air-water heat angers and is then drawn back towards the rotor via the top bottom return paths. re 1995, axial ventilation had been used in the generators. In , the top and bottom rotor cover plates were removed to ide some radial ventilation and reduce stator winding eratures. This resulted in a 15% increase in the total air flow and an average 5ºC decrease in the temperatures of the ings. However, an additional modification was still needed rther reduce stator winding temperatures and extend the

ational life of the generators.

Figure 2. A radial cross-section of a Roxburgh generator with arrows indicating the air flow paths in the radial-axial ventilation system. Experiments

Experiments were carried out to measure velocities and pressures, and to visualise flow patterns, at critical points within the generators. Qualitative and quantitative results from the experiments were needed to develop and validate the computational fluid dynamics (CFD) models used to redesign the fan blades. Experiments were essential because of the limited understanding of the ventilation system and the small number of permanent sensors. The integration method as outlined in [2] was used to measure the air flow rate through the stator, top and bottom return paths and an entire generator. In these experiments the cross-sectional area of interest was divided into a number of sub-areas. A time-averaging turbine anemometer was then used to measure the mean velocity through each sub-area which enabled the calculation of the air flow rate through the cross-section. Using this method the following results were obtained: • The total air flow rate through a generator is 23m3/s as

measured across four heat exchangers and then doubled. • There is a 70%/30% imbalance in air flow in favour of the top

return path. • Almost all (95%) of the air flows past the stator windings. The total air flow rate of 23m3/s was confirmed using an energy balance across a heat exchanger. Two conclusions can be drawn from the above results: • The speed, not the proportion, of air flow past the stator

windings must be increased for improved cooling. • The imbalance between the top and bottom air flows shifts the

uneven flow distribution due to axial ventilation towards the bottom of the stator.

Smoke was used to visualise the flow patterns in the top of the generators. The main ventilation mechanism for the generators was identified by piecing together these flow patterns. From the smoke it was noticeable that the largest proportion of the air flowing past the stator windings still came from axial ventilation. High-pressure is created at the ends of the rotor poles and around the end-windings by the centrifugal pumping of the fan blades. The high-pressure drives air into the low-pressure region between the rotor poles. This ventilation mechanism caused large amounts of very turbulent swirling air around the fan blades which was a significant source of power loss. Static pressure measurements before and after the fan blades showed that there was a 100Pa increase across them from centrifugal pumping. A five-hole probe of the forward facing pyramid type was used to measure velocities at critical points in observed air flow patterns.

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g the five-hole probe the speed, yaw angle and pitch angle of rtant air flows were calculated from five pressure readings.

canivalve was used to index each reading which was then erted to a voltage by a pressure transducer. As the ivalve operation of a five-hole probe requires steady flow, it impossible to develop a comprehensive velocity map for the of the generators. The most important measurement made the probe was the mean velocity of the swirling air around an blades which was 20m/s.

tion Strategy

d upon the experimental results the redesign and cement of the fan blades on the rotor rim was chosen as the solution. There are forty-four poles and fan blades on a rotor one blade between every pair of poles alternating between nd bottom. Figure 3 illustrates the form and position of the nal fan blades. The redesign and replacement of the fan es was chosen for the following reasons: is less expensive than other solution strategies such as esigning and replacing the air coolers. e fan blades act on a large proportion of the air flow. placement of the fan blades is simple to implement. ere is scope for significant improvement in the original fan de design. e number of fan blades can be adjusted to balance air flow rease against windage loss.

e 3. The form and position of an original fan blade on the top ce of the rotor. There is one fan blade between every pair of rotor alternating between top and bottom.

Modelling

CFD models were meshed using Gambit [3] and then run in ent solver [4]. The main alternative to using CFD modelling trial and error testing of different fan blade designs.

ever, trial and error testing would have been much more nsive because a generator would need to spend a lot of time ne. Like most practical CFD problems a suitable compromise een computational effort and the level of resolution of flow res had to be made for the models. The efficient use of mesh important because the CFD models were used as a design to determine the best form for the fan blades. The major vantage of CFD modelling was that a number of

oximations and assumptions had to be made to simplify the plex air flow and geometry of the generators. Without these lifications the problem would be impractical to solve given omputer resources available. The main simplifications used e CFD models were: t modelling heat transfer. If heat transfer were modelled the el of boundary layer resolution would make the problem too e consuming to solve. It was assumed that increasing the air w past the stator windings would reduce their temperature.

• Assuming the top and bottom air flows to be identical despite the measured imbalance. Computational effort was therefore halved by only considering the top of the generators. In terms of fan blade design the main consequence of this simplification was not accounting for the effect of alternating top and bottom blades.

• Using porous media to simulate the effects of small-scale geometry and reduce the size of the mesh. In the Fluent solver a fluid zone was defined as porous media and a pressure gradient was set to match the known gradient across the actual geometry. For example the stator is an iron annulus with many radial ducts and can be modelled as a fluid zone with infinite resistance in the axial and circumferential directions, but with a finite resistance giving the actual pressure drop in the radial direction.

Despite these simplifications it was assumed that relative differences between models would be repeated, to some extent, in reality. Although closed-circuit self-driven CFD models would better represent the ventilation system they would not provide the level of centrifugal pumping that was measured in the experiments. This is because of their simplified geometry and limited resolution of important flow features. Instead open-circuit models were used with velocity inlet and mass flow outlet boundary conditions to focus in on the fan blades. The inlet velocity was derived from the measured total air flow rate and a component of swirl was imparted to this flow by the rotor. Two-Dimensional Axisymmetric Model

A radial cross-section through the top half of the generators was used as the geometry of the 2D axisymmetric model. The geometry, contours of velocity magnitude and velocity magnitude vectors at inlet and outlet are shown in Figure 4. Since this model was axisymmetric, it accounted for the circumferential flow caused by the rotation of the rotor. Modelling circumferential flow was important because of the high degree of coupling between rotating flow and the radial flow which drives the air through the generators. The main purpose of this model was to determine the porous media settings needed to obtain 25% of the inlet air flow through the rotor. Since this air flow could not be measured with a simple experiment, the choice of 25% was an estimate based upon the amount of smoke drawn into the rotor. Because of its geometrical limitations, the 2D model was developed to the stage of getting realistic flow patterns similar to those observed during the experiments while still allowing 25% of the inlet air flow through the rotor. Circumferential flow was set inside the rotor and prevented in the radial ducts of the stator using the user defined function capability of the Fluent solver.

Figure 4. The geometry, contours of velocity magnitude and velocity magnitude vectors at inlet and outlet for the 2D axisymmetric model.

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e-Dimensional Models

22nd wedge of the top half of a generator was the smallest dic unit and was used as the geometry of the 3D models to mise the size of the domain. To achieve this both periodic symmetry boundary conditions were used in the Fluent r. The geometry, contours of velocity magnitude for the and velocity magnitude vectors at inlet and in the stator are n in Figure 5. In the 3D models the flow inboard of the r was solved in a rotating reference frame that moves with otor. The remainder of the domain was solved in a stationary ence frame. The 3D models simultaneously solved for steady in both reference frames. This is because using porous ia instead of physical blockage meant that the flow was pendent of the position of the rotor. To reduce the amount of in the 3D models the air flow through the rotor was set the results of the 2D model and was not determined from est of the solution. A close qualitative agreement with the rimental results was obtained for the model of the existing low. However, the model under predicted the pressure boost ided by the rotor compared with that measured in the riments because of the practical limit on the amount of dary layer resolution. As a result all of the pressures

nstream of the rotor were lower than their measured values. e relative changes in pressure and velocity were used in signing the fan blades, close experimental and model values not critical.

e 5. The geometry, contours of velocity magnitude for the rotor and ity magnitude vectors at inlet and in the stator of the 3D model.

deficiencies with the original fan blades were highlighted by model and are shown by relative velocity magnitude vectors nd the blades in Figure 6. The flow patterns in Figure 6 are ative of those in the immediate vicinity of the fan blades. e fan blades are designed to pump air in an outward radial tion to create high-pressure at the ends of the rotor poles. ever, a significant proportion of the air acted on by the es is not driven in an outward radial direction or between the poles and is wasted. The inefficient performance of the fan es is compounded by their positioning. This is because air is n from between the rotor poles by the low-pressure on the on side of the blades. fan blade was removed from the 3D model to investigate the rtance of the blades in the ventilation system. This model ated that 70% of the pumping power transferred from the lades and the outer surface of the rotor to the surrounding as attributable to the blades. As a result the fan blades are nsible for creating most of the high-pressure at the ends of otor poles and therefore determine the performance of the ventilation component.

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Figure 6. Relative velocity magnitude vectors on the pressure face, rotor pole gap and neighbouring rotor surface for the original fan blades. Redesigning the Fan Blades

The 3D model was used as a design tool to determine the most suitable form for the new fan blades. Since the inlet air flow rate for each fan blade design was unknown, the inlet boundary conditions for the original blades were used in all of the cases. It was assumed that a constant flow would not have a significant effect on the relative performance of the blades. Both the pressure boost across each fan blade and the air flow rate driven between the rotor poles were used to indicate the relative performance of each design. Since almost all of the air flow was measured to travel through the stator, the aim of each fan blade design was to increase the flow rate past the stator windings by increasing the air speed. There were two general approaches for redesigning the blades: • Improve the radial pumping of the fan blades to increase the

high-pressure at the ends of the rotor poles. • Use scooping fan blades to deflect more air between the poles. Although a combination of both approaches was used, there was a focus on deflecting the air because it was a more direct method and avoided adding to the large amount of swirling air around the fan blades. This swirling air decreases the apparent flow speed onto the blades and therefore their performance. The following features were incorporated into the design of the new fan blades: • Scooping curvature to deflect more air between the rotor poles. • Parallel alignment with a slight offset from the edge of each

downstream rotor pole. This is to maximise the cross-sections of the rotor pole gaps for deflected air while minimising leakage on the low-pressure side of the fan blades.

• Forward curvature of the leading edge to enhance the flow onto the fan blades.

• Lengthening the fan blades to increase their radius ratio and therefore their pumping effect.

To evaluate the performance of the redesigned fan blades the pressure boost, local windage and viscous losses for the blades were calculated relative to the original ones. From these calculations it was found that the redesigned fan blades provide a 40% greater pressure boost, but cause a 10% (about 3kW) local increase in windage losses which was low compared with the overall windage for each generator of 200 kW. The viscous losses for the redesigned and original fan blades were calculated as a proportion of the pumping power from the blades and the outer surface of the rotor. For the new fan blades this viscous loss was calculated as 40% compared with 60% for the old blades, showing the improved performance of the new fan blades. The redesigned fan blades were also intended to provide sufficient outward radial air flow to at least maintain the temperature of the stator end-windings. Figure 7 illustrates the basic form and flow patterns for the new fan blades.

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e 7. Relative velocity magnitude vectors for the pressure face, rotor gap and neighbouring rotor surface for the redesigned fan blades.

actical implementation of the chosen fan blade design from CFD models was developed to accommodate simple

ufacture, strength requirements and the existing bolting gement. A prototype of the redesigned fan blades was made

is shown next to an original blade in Figure 8.

e 8. A prototype of the redesigned fan blades beside an original one.

clusion

modelling was able to be used as a cost effective design tool velop a practical fan blade design up to the prototype stage

nstallation on the Roxburgh generators. This CFD modelling ates that the redesigned fan blades will increase the air flow the stator windings and therefore lower their temperature. basic form of the fan blades is also applicable to other o-generators using axial ventilation. However, the actual rmance of the fan blades will not be known until they are

in practice.

nowledgments

authors wish to thank the Foundation for Research, Science Technology (New Zealand), under contract number

X9901, for their assistance during the course of this project.

rences

orgna, H. & Garcia, A., Achieving Maximum Performance rom Hydro-Generator Ventilation Systems, International ournal on Hydropower & Dams, Aqua-Media International, ssue 3, 1997, 84-87. ilipan, V., Budin, R. & Mihelic-Bogdanic, A., Air Flow easurement on Hydro-Generators, Water Power & Dam onstruction, Surrey, Reed Business Publishers, January 993, 44-46. ambit 1.3, Fluent Inc., 1999. luent 5.4, Fluent Inc., 1999.

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