Research Papers: Gas Turbines: Turbomachinery

Modeling and Experimental Investigation of an Oil-Free Microcompressor-Turbine Unit for an Organic Rankine Cycle Driven Heat Pump

[+] Author and Article Information
Jonathan Demierre

Industrial Energy Systems Laboratory (LENI),
Ecole Polytechnique Fédérale de
Lausanne (EPFL),
Lausanne 1015, Switzerland
e-mail: jonathan.demierre@a3.epfl.ch

Antonio Rubino

Laboratory for Applied Mechanical
Design (LAMD),
Ecole Polytechnique Fédérale de
Lausanne (EPFL),
Neuchâtel 2002, Switzerland
e-mail: antonio.rubino@epfl.ch

Jürg Schiffmann

Laboratory for Applied Mechanical
Design (LAMD),
Ecole Polytechnique Fédérale de
Lausanne (EPFL),
Neuchâtel 2002, Switzerland
e-mail: jurg.schiffmann@epfl.ch

1Corresponding author.

Contributed by the Turbomachinery Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received July 9, 2014; final manuscript received July 28, 2014; published online October 7, 2014. Editor: David Wisler.

J. Eng. Gas Turbines Power 137(3), 032602 (Oct 07, 2014) (10 pages) Paper No: GTP-14-1347; doi: 10.1115/1.4028391 History: Received July 09, 2014; Revised July 28, 2014

Domestic heating and cooling will more and more have to rely on heat pumps (HPs) in order to support a more rational use of primary energy consumption. The HP market is mainly dominated by electrically driven vapor compression cycles and by thermally driven sorption processes. The drawback of electrically driven vapor compression cycle is their dependence on an electrical grid and the fact that they increase the winter or summer electricity peak demands. Hence, a thermally driven vapor compression cycle would offer substantial advantages and flexibility to the end user for heating and cooling applications. This paper presents the investigation of an oil-free compressor-turbine unit (CTU) used for a thermally driven HP (TDHP) based on the combination of a HP compression cycle and an organic Rankine cycle (ORC). The CTU consists of a radial inflow turbine and a centrifugal compressor of the order of 2 kW each, directly coupled through a shaft supported on gas lubricated bearings. The CTU has been tested at rotor speeds in excess of 200 krpm, reaching compressor and turbine pressure ratios up to 2.8 and 4.4, respectively, and isentropic efficiencies around 70%. Comparisons between the experimental data and predictions of models, that are briefly described here, have been carried out. A sensitivity analysis based on the experimentally validated models shows that tip clearance, for both compressor and turbine, and surface roughness of the compressor are key parameters for further improving performance.

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Fig. 1

Grassmann diagram of the ORC-HP exergy flows. Results corresponding to experimental run #12 (see Table 3). Environmental conditions: P0 = 1 bar and T0 = 10 °C.

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Fig. 2

Compressor stage geometry with no inlet guide vanes and vaneless diffuser [14]

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Fig. 3

Radial inflow turbine stage geometry

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Fig. 4

Comparison between measured and predicted rotor windage losses

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Fig. 5

Experimental setup diagram

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Fig. 6

Static pressures at turbine rotor inlet (PT,4), at compressor impeller exhaust (PC,4), at compressor inlet (PC,in), and in the bearing housing (PB)

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Fig. 7

Relative leakage mass flow rate out of the CTU as a function of the ratio between pressure at turbine rotor inlet (PT,4) and pressure in the bearing housing (PB)

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Fig. 8

Predicted and measured turbine mass flow rate for different turbine pressure ratios

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Fig. 9

Predicted and measured turbine isentropic efficiency for different turbine pressure ratios

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Fig. 10

Predicted and measured compressor isentropic efficiency for different compressor pressure ratios

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Fig. 11

Sensitivity analysis: (a) CTU efficiency as a function of the compressor pressure ratio and (b) mean ∂ηCTU/∂ɛt as a function of the mean compressor pressure ratio

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Fig. 12

Sensitivity analysis: (a) CTU efficiency as a function of the compressor pressure ratio and (b) mean (∂ηCTU/∂k+) as a function of the mean compressor pressure ratio

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Fig. 13

Sensitivity analysis of the dimensionless tip clearance: (a) turbine pressure ratio versus compressor pressure ratio and (b) turbine power versus compressor pressure ratio

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Fig. 14

Sensitivity analysis of the dimensionless surface roughness: (a) turbine pressure ratio versus compressor pressure ratio and (b) turbine power versus compressor pressure ratio




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