AIAA-2009-4835

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American Institute of Aeronautics and Astronautics2

= isentropic efficiency

t = expansion ratio of the vaneless counter-rotating turbine

I. Introduction

Vaneless Counter-Rotating Turbine (VCRT) is composed of a highly loaded single stage high pressure

turbine (HPT) and a single stage vaneless counter-rotating low pressure turbine/rotor (LPT/LPR). Comparing

with the conventional two stage turbine, the VCRT can offer significant benefits due to the parts elimination and

size reduction. These benefits include elevating thrust-to-weight ratio, reducing cooling flow and cost, and so on.

From the 1950s, counter-rotating turbines have been carefully investigated. Wintucky et al. (1958, [1]) analyzed

the effects of loading coefficient, turbine exit whirl and rotational speed ratio on the efficiency of the counter-

rotating turbine. Louis (1985, [2]) carried out an investigation on two counter-rotating turbines with different

configurations. The results indicate that the efficiencies of these two counter-rotating turbines are higher than

conventional turbines with the same stage loading coefficient. In other words, the counter-rotating turbines have

higher stage loading coefficients compared with conventional turbines with the same efficiency. The work

performed by Sotsenko and Ponomariov (1990, [3, 4], 1992, [5]) illustrates the potential of counter-rotating

turbine. Compared with conventional turbine, the investigation results indicate that the counter-rotating

configuration can obtain 5% increased pressure ratio of the compressor, 23% shorten engine length, 35-40%

decreased airfoil number of the gas turbine and 45-50% decreased airfoil number of the power turbine at the same

fuel consumption. Weaver et al. (2000, [6]) focused on understanding the physical parameters influencing the

unsteady forces causing the blade excitation in a transonic vaneless counter-rotating turbine. In this investigation,

the blade surface unsteady pressures were measured and analyzed. Using these measured data, the proper CFD

modeling requirements to engineering design assessments were determined. Haldeman et al. (2000, [7])

investigated the pressure loading on a vaneless counter-rotating turbine by means of a short-duration shock tunnel

facility and CFD tool. The results indicate that the CFD code can qualitatively capture the flowfield physics.

However, some additional calibration on the code needs to be performed in order to fully match experimental data

quantitatively. Keith et al. (2000, [8]) introduced the tests on the Controlled Pressure Ratio Engine (COPE) turbinesystem. The turbine system consists of a single stage high pressure turbine and a two stage vaneless counter-

rotating low pressure turbine. The experimental results show that the performance levels of the HPT meets the pre-

test expectations, and the LPT also meets performance objectives across the tested range. Zhao et al. (2007, [9],

2008, [10], 2009, [11, 12]) performed numerical and experimental investigations on unsteady flow characteristics

and inlet hot streak effects in a vaneless counter-rotating turbine. In these investigations, the unsteady pressure

fluctuations on the surface of airfoils and tip region of the HPT rotor were captured. And the effects of

temperature ratio and tip leakage flow on hot streak migration also were obtained. These results will be used into

the unsteady design of vaneless counter-rotating turbine in the future.

In this paper, the flow characteristics of a vaneless counter-rotating turbine at off-design conditions will be

explored by means of a blow-down short duration turbine test facility and a three-dimensional multiblade row

steady Navier-Stokes code.

II. Experimental Facility

The IET (Institute of Engineering Thermophysics, Chinese Academy of Sciences) blow-down turbine facility is

a transient wind tunnel, which can be used to simulate flow conditions for most modern high pressure axial

turbines. The blow-down short duration turbine facility is able to substitute a continuous turbine facility in a

majority of turbine testing on flow and heat transfer measurements. The valuable test time in the IET facility is

about 300-500 milliseconds. This test time is sufficiently long compared to the flow and heat transfer

characteristic time in a high speed turbine stage. So, the turbine operates in a quasi-steady state during test process.

The blow-down short duration turbine facility has some advantages compared to the continuous long duration

facility. One advantage of the short duration facility is the lower cost of construction, operation and maintenance.

A

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The other is more convenient measurement of heat transfer characteristic. The metal temperature is the same as

room temperature because the test time is short compared to the thermal transient of the turbine blade. An almost

constant gas-to-metal temperature ratio is maintained during the test time. The relatively favorable environment is

suitable for the application of the heat flux gauges. Then, the heat transfer data are readily acquired in the short

duration test.

The schematic of the blow-down short duration turbine facility is given in Figure 1. Major components shown

from upstream to downstream are the supply tank (12 m3), fast response valve, test section, tail cone and vacuum

tank (20 m3). Cross-sectional view of the VCRT internal flow path is shown in Figure 2.

Supply Tank

Fast Response Valve Test Section Tail Cone

Vacuum Tank

Figure 1. Schematic of the IET blow-down short duration turbine facility.

HPT Stator HPT Rotor LPT Rotor

Figure 2. Cross-sectional view of the VCRT internal flow path.

III. Numerical Algorithm

NUMECA software systems are employed to study this problem. The numerical method is described in details

in the user manual (2005, [13]). Here only a brief description about the main features is reported.

The governing equations in NUMECA are the time dependent, three-dimensional Reynolds-averaged Navier-Stokes equations. The solver of NUMECA is FINE/Turbo and it is based on a cell centered finite volume

approach, associated with a central space discretization scheme together with an explicit four-stage Runge-Kutta

time integration method.

Residual smoothing, local time-stepping, and multi-gridding are employed to speed up convergence to the

steady state solution.

Various turbulence models have been included in the solver for the closure of governing equations. The widely

used approach based on one transport equation (Spalart and Allmaras, 1992, [14]) has been selected in this paper.

The Spalart-Allmaras model has become quite popular in the last years because of its robustness and its ability to

treat complex flows. The main advantage of Spalart-Allmaras model when compared to the one of Baldwin-

Lomax is that the turbulent eddy viscosity field is always continuous. Its advantage over the k- model is mainly

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its robustness and the lower additional CPU and Memory usage.

The time step limitations arising from the turbulent source terms are accounted for in the model so that

computations can be performed with the maximum allowable CFL number without penalizing numerical

efficiency.

IV. Boundary Conditions

The theory of characteristics is used to determine the boundary conditions at the inlet and exit of computational

domain. At the inlet, total pressure, total temperature and circumferential and radial flow angles are specified as

many constants in the simulation. Due to selecting the Spalart-Allmaras turbulence model, the kinematic turbulent

viscosity should be specified in the inlet boundary conditions. In this paper, it is 0.0001m2/s.

At the exit, the circumferential and radial velocity components, entropy and the downstream running Riemann

invariant are extrapolated from the interior of the computational domain. The static pressure, P6, is specified at the

hub of the exit and the static pressure values at all other radial locations are obtained by integrating the equation

for radial equilibrium. Periodicity is enforced along the outer boundaries of H-O-H grids in the circumferential

direction.

No-slip boundary conditions should be enforced at solid wall surfaces for viscous simulations. In this paper,

absolute no-slip boundary conditions are enforced at the hub and tip end walls of the HPT stator regions, along the

surface of the HPT vane, and along the casing walls of the HPT rotor and LPR regions. Relative no-slip boundary

conditions are imposed at the hub end walls of the HPT rotor and LPR regions, and the surfaces of the HPT rotor

and LPR blades. It is assumed that the normal derivative of pressure is zero at the solid wall surfaces, and that the

walls are adiabatic.

V. Vaneless Counter-Rotating Turbine

The VCRT studied in this paper is composed of a highly loaded single stage HPT coupled with a vaneless

counter-rotating LPT/LPR. It has high expansion ratio and operates in transonic regimes. The design conditions of

the VCRT are shown in Table 1.

Table 1. The flow conditions in the VCRT

Inlet total temperature (K) 500

Inlet total pressure (kPa) 300

Mass flow (kg/s) 17.7

Rotational speed of HPT rotor (RPM) 7162

Rotational speed of LPR (RPM) -6778

Expansion ratio of HPT 2.93

Expansion ratio of LPT 2.07

SWR 1.77

For these calculations, no tip clearance of the rotor is modeled. And the typical y+ values of less than 15 are

used at the boundaries. The three-dimensional H-O-H grid topologies are showed in Figures 3 to 5.

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2 3 4 5 6 7100

200

300

400

500

600

nrs=0.41 (Experimental Results)

nrs

=0.41 (Numerical Results)


nrs=0.60 (Numerical Results)





N / k W

t

Figure 6. Power characteristic curves of HPT.

Figure 7 shows the experimental and numerical power characteristic curves on the LPT. The numerical code

qualitatively predicts the power performance of the LPT. The power performance of the LPT meets the design

requirement at full operating range. The results in Fig. 7 indicate that the increase of the rotor rotation speed tends

to decrease the power of the LPT. The comparison between HPT data and LPT data illustrates that there is a more

optimized rotation speed ratio for the VCRT. In this paper, the optimum rotation speed ratio of the VCRT is

closed to 1.8. The results also indicate that the power performance of the LPT still has a potential to be improved.

2 3 4 5 6 70

100

200

300

400n

rs=0.38 (Experimental Results)


N / k W

t

2 3 4 5 6 70

100

200

300

400 nrs=0.57 (Experimental Results)


N / k W

t

0.38rsn 0.57rsn

2 3 4 5 6 70

100

200

300

400n



N / k W

t

2 3 4 5 6 70

100

200

300

400n



N

/ k W

t

0.73rsn 0.94rsn

Figure 7. Power characteristic curves of LPT.

Figure 8 shows the experimental and numerical power characteristic curves on the VCRT. The power

performance meets the design objective. The numerical tool has qualitative prediction ability. The results indicate

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that the increase of the rotor rotation speed tends to increase the power of the VCRT. And the effect of the rotor

rotation speed on the power of VCRT becomes weaker at higher rotation speed region.

2 3 4 5 6 7200

400

600

800

1000n









N / k W

t

Figure 8. Power characteristic curves of VCRT.

The SWR characteristic curves of the VCRT are shown in Fig. 9. The numerical data are very agreement with

the experimental data. The prediction accuracy of this code is verified. It is competent for this investigation. The

results in Fig. 9 show that the increase of the rotor rotation speed tends to increase the SWR of the VCRT. And

when the rotor rotation speed is increased, the operation range of the VCRT is reduced.

2 3 4 5 6 70

1

2

3

4

5n









R W

t

Figure 9. SWR characteristic curves.

Figure 10 shows the experimental and numerical isentropic efficiency characteristic curves on the VCRT. Theresults show that the efficiency of the VCRT is increased with the increase of the rotor rotation speed. The

predicted results are well agreement with experimental data. It illustrates the numerical code can qualitatively

predict the efficiency of the VCRT.

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2 3 4 5 6 750

60

70

80

90

100



t

2 3 4 5 6 750

60

70

80

90

100n



t

0.40rsn 0.59rsn

2 3 4 5 6 750

60

70

80

90

100n



t

2 3 4 5 6 750

60

70

80

90

100n



t

0.77rs

n 0.97rsn

Figure 10. Isentropic efficiency characteristic curves of VCRT.

VII. Conclusions

In this paper, the flow characteristics of a vaneless counter-rotating turbine at off-design conditions areinvestigated by means of a blow-down short duration turbine test facility and a three-dimensional Navier-Stokes

code. About 100 operating modes are simulated in this investigation. Depending on the experimental and

numerical data, the performance curves of the vaneless counter-rotating turbine are obtained. The qualitative

prediction ability of the numerical code is also verified. The following main conclusions have been drawn:

1) The increase of the rotor rotation speed tends to increase the ratio of specific work of the high pressure

turbine to that of the low pressure turbine and the efficiency of the vaneless counter-rotating turbine.

2) When the rotation speed of the rotor increases, the specific work of the low pressure turbine is decreased, and

the effective operation range of the vaneless counter-rotating turbine is reduced.

Acknowledgments

This work is supported by the Award Fund of the President of CAS. The support of the Wu Chung Hua AwardFoundation is gratefully acknowledged.

References1Wintucky, W.T., and Stewart, W.L., “Analysis of Two-Stage Counter-Rotating Turbine Efficiencies in terms of Work and

Speed Requirements,” NACA RM E57L05, 1958.2 Louis, J.F., “Axial Flow Contra-Rotating Turbines,” ASME Paper 85-GT-218, 1985.3 Sotsenko, Y.V., “Thermogasdynamic Effects of the Engine Turbines with the Countra-Rotating Rotors,” ASME Paper

90-GT-63, 1990.4 Ponomariov, B.A., “New Generation of the Small Turboshaft and Turboprop Engines in the USSR,” ASME Paper 90-

GT-195, 1990.

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