OPEN-LOOP AERODYNAMIC PERFORMANCE TESTING OF A 105,000 RPM OIL- FREE COMPRESSOR-EXPANDER FOR SUBSURFACE NATURAL GAS COMPRESSION AND REINJECTION

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1 Proceedings of ASME Turbo Expo 202 GT202 June -5, 202, Copenhagen, Denmark DRAFT GT OPEN-LOOP AERODYNAMIC PERFORMANCE TESTING OF A 05,000 RPM OIL- FREE COMPRESSOR-EXPANDER FOR SUBSURFACE NATURAL GAS COMPRESSION AND REINJECTION Aaron M. Rimpel Southwest Research Institute San Antonio, TX, USA Timothy C. Allison Southwest Research Institute San Antonio, TX, USA J. Jeffrey Moore Southwest Research Institute San Antonio, TX, USA Joseph S. Grieco Halliburton Energy Services Carrolton, TX, USA Perry C. Shy Halliburton Energy Services Carrolton, TX, USA John M. Klein Rotor-Therm Anchorage, AK, USA Jerry L. Brady BP Exploration Alaska Anchorage, AK, USA ABSTRACT An aerodynamic performance test stand has been developed for validation of the performance of a 05,000 rpm compressor-expander which is intended for subsurface natural gas reinjection. The turbomachine consists of a two-stage centrifugal compressor, which is driven by a single-stage expansion turbine. The rotor is supported by foil gas journal bearings and a spiral-groove gas thrust bearing. The test stand is configured for open-loop testing of the overall compressor and turbine performance with air as the working fluid and atmospheric pressure at the compressor suction and turbine discharge locations. Several performance curves were generated for each component ranging from 73,500-5,500 rpm (70-0% of design speed). In general, measured compressor head was slightly lower than predictions, while measured efficiencies were close to predicted values. The turbine had higher flow than predicted, due in part to a larger flow area in the turbine. The turbomachine has shown acceptable performance on the open-loop test stand, and further testing at higher-pressure closed-loop conditions are planned. Keywords: Rotordynamics, micro-turbomachinery, oilfree, foil bearing, centrifugal compressor, expander / expansion turbine / turbo-expander, hole-pattern seal, labyrinth seal, stability analysis, seal taper sensitivity analysis INTRODUCTION. Background The hydrocarbon recovery in some oil fields is limited due to the gas handling restrictions on the surface. The high cost of adding facilities to increase the gas handling capacity has led to the development of a tool that will process, compress, and reinject up to 30% of the gas in a high gas-to-oil ratio (GOR) well (greater than 5,000 standard cubic feet of gas per stock tank barrel of oil). This technique allows additional black oil to be produced due to the increased gas rate production, and the gas condensate production is able to be increased as the recycled gas will carry the condensates to the production stream. Brady et al. [] discusses the initial investigation and validation for the Subsurface Process And Reinjection Compressor (SPARC) tool development. The main benefit of the SPARC tool is the ability to provide additional gas handling capacity and maintain reservoir pressure to existing production fields at a fraction of the cost of a surface-based operation..2 Description of SPARC Tool Figure represents the SPARC concept presented by Brady et al. [] and will be used to describe the tool s basic functions. The SPARC device includes a three phase separator at the bottom of the tool which will prevent most liquids and solids from passing through the turbo-expander or turbine (separator Copyright 202 by ASME

2 is labeled as pre-swirl auger in Figure ). The remaining solids and liquids bypass around the turbomachinery section. The energy in this production stream is usually wasted by passing through a series of chokes within the production system. A turbine bypass valve (not shown) allows the gas flow to enter the turbine only after the well has been cleaned up and is free from debris usually seen while bringing a shut-in well back online. Until these conditions are achieved, the valve remains closed and flow bypasses the turbomachinery section entirely. Reservoir gas is filtered (not shown) and used to cool the oilfree bearings. The production stream is expanded through the turbine, rejoins with the bypass flow, and is separated once again by an auger separator the gas is directed to the compressor suction while the other solids/liquids continue to the surface. A recycle valve is utilized to prevent the compressor from operating in surge conditions. The gas which exits the compressor is finally reinjected back into the formation. The following sections describe the turbomachinery components and the test rig in more detail. Figure : Schematic of SPARC concept for functional description (adapted from []). 2 DESCRIPTION OF THE TURBOMACHINERY SECTION The turbine and compressor are mounted on the same shaft in a double-overhung arrangement with respect to the bearings. The turbine is a single-stage, shrouded, mixed-flow turbine. Note that the early concept in Figure shows a radial turbine, but a mixed-flow turbine allows for a smaller-diameter design compared to a conventional radial-inflow turbine with the same flow capacity. The compressor is a two-stage centrifugal compressor which incorporates shrouded impellers with threedimensional blades and splitter vanes. Splitter vanes allow for an appropriate blade solidity (ratio of blade chord length to pitch) while providing a larger throat area so that more mass flow may pass before choking occurs. Both diffusers are vaned for improved efficiency and utilize Low Solidity Diffuser (LSD) vanes. LSD vanes improve pressure recovery and efficiency of centrifugal compressors, but unlike full diffuser vane designs, flow range is not hindered. The design speed of the turbine and compressor is 05,000 rpm (05 krpm), and the design compressor power consumption at down hole conditions is approximately 460 hp (343 kw). Given that the turbine weighs approximately lbf (0.907 N) and the combined compressor impeller weight is approximately 8 lbf (0.703 N), the power-to-weight ratio of the SPARC is an order-ofmagnitude larger than a typical commercial centrifugal impeller. Both the radial bearings and the thrust bearing are fluid film bearings that operate on the filtered process gas as mentioned in the previous section. The radial bearings are.50 in (38. mm) journal diameter, Generation III foil gas bearings with length-diameter ratio of unity. The foil gas bearings are categorized as Generation III due to stiffness and damping characteristics which vary both axially and circumferentially [2]. The thrust bearing is a smooth rotor, spiral groove stator gas thrust bearing. Hole-pattern stator seals were designed for down hole gas conditions and were initially used for aerodynamic performance testing in the open-loop test stand. At down hole gas conditions, hole-pattern seal stiffness is significant (on the same order of magnitude as the bearings), and the lowest rotor natural frequency is expected to be near 35 kcpm (0.33 of synchronous frequency at design speed). However, at openloop air test conditions, hole-pattern seal stiffness is significantly lower (two orders of magnitude lower than the bearings), and the lowest rotor natural frequency occurs near 2 Copyright 202 by ASME

3 3-5 kcpm ( of synchronous frequency at design speed), which is below the predicted cross-over frequency of the hole-pattern seals. (Note that in hole-pattern seal analysis, the cross-over frequency is the frequency where effective damping of the seal transitions from negative to positive. Potential instability of the rotor-bearing-seal system can result if the first natural frequency is below the cross-over frequency of the hole-pattern seals, especially if the seals exist at the antinodes of the corresponding mode shape. Additional discussion on this behavior can be found in Ref. [3] or in the companion paper.) Instability at this excitation frequency was observed during tests as the rotor speed approached 00 krpm, preventing full-speed aerodynamic performance testing with hole-pattern seals. As a solution, labyrinth seals which have better stability characteristics than hole-pattern seals at the test conditions were also used to complete aerodynamic performance testing, and performance data was able to be gathered at speeds up to 5.5 krpm (0% of design speed). A companion paper presents the rotordynamic investigation of the open-loop aerodynamics test stand with the different seal configurations. 3 DESCRIPTION OF OPEN-LOOP AERODYNAMICS TEST STAND An open-loop test stand was constructed to test the overall compressor and turbine performance with air as the working fluid and atmospheric pressure at the compressor suction and turbine discharge locations. The test stand utilizes most of the turbomachinery components designed for the actual tool, but a different housing was utilized to permit sufficient instrumentation. Due to the small size of the test rig, it was not physically possible to strictly follow the ASME PTC-0 guideline [4], but measurement redundancy was employed at all critical performance measurement locations. Figure 2 shows a photograph of several of the test rig components. Figure 3 shows the schematic of the open-loop test stand. The numbers in the schematic indicate various measurement station locations. There are three main open-loops (indicating venting to atmosphere instead of rejoining the source) in the test stand: the compressor loop, the turbine loop, and the bearing cooling loop. The compressor loop draws in ambient air and discharges to atmosphere through a throttle valve. The turbine loop source has a throttle valve leading into the turbine, and the discharge after expansion is to atmosphere. With foil bearings, hydrodynamic pressure must be generated to produce a physical clearance i.e., when the journal is not rotating, the bearing surface (top foils) are preloaded against the journal. While coatings are typically employed to mitigate wear and excessive dry friction during starts/stops, there is still a relatively high level of torque required to overcome the static friction in a start-up case when compared to the full-hydrodynamic drag torque. Burnishing of the coating through multiple starts/stops typically results in reduction of break-away torque magnitude. Therefore, as shown in Figure 3, a supply of higher pressure air (200 psig 3.8 bar) was occasionally used for the start-up of the rotor depending on the level of radial bearing break-in. Once several starts/stops were exercised on the bearings, shop air supply (00 psig 6.9 bar) was sufficient to overcome the static break-away torque of the bearings. The bearing cooling supply is from the same compressed air source as the turbine and also discharges to atmosphere. Static pressures and flow rates are monitored at the inlet locations of the housing (stations 6.00 and 6.0); static pressure is monitored at the discharge/vent location (station 6.20) A G H I B in mm F E D C Figure 2: Photograph of test rig components: (A) test rig housing assembly, (B) large labyrinth seal, (C) compressor balance piston hole-pattern seal, (D) stage 2 compressor impeller, (E) stage compressor impeller, (F) interstage diaphragm section, (G) stage 2 diffuser and de-swirl cascade, (H) turbine hub hole-pattern seal, and (I) complete rotor assembly. 3 Copyright 202 by ASME

4 Shop air Rotameter Accumulator for startup, 200 psig supply Ambient air.0.03 Venturi flow meter Rotameter Control valve Check valve Orifice plate flow meter Shop air 5.0 Compressor Control valve Turbine Bearing Leakage Figure 3: Open-loop aerodynamic performance test stand schematic. Numbers indicate measurement station locations. For the compressor, flow rate is measured with a venturi flow meter, and static temperature and static pressure are measured prior to entering the stage impeller. Static temperature and static pressure are also measured at the compressor exit downstream of the stage 2 impeller exit, which is after the diffuser and the deswirl cascade in the discharge annulus. A variety of static pressure taps are also placed at multiple locations along the flow path through the compressor stages where physically possible. Table lists the various measurements for compressor performance testing, and Figure 4 depicts a cross-section of the test rig which illustrates the measurement locations at the compressor end. Table : Compressor measurement location descriptions. Station Physical Measurement No. Description Type.00 Ambient conditions (global) Relative humidity*.0 Ambient conditions (local, venturi inlet) Temperature.03 Venturi flow meter Mass flow rate (calculated)*.05 Venturi exit Total pressure*.08 Stage impeller inlet Static temperature*, static pressure.35 Stage diffuser vane channel Static pressure 2.25 Stage 2 impeller exit Static pressure 2.29 Stage 2 diffuser vane inlet Static pressure 2.60 Stage 2 deswirl vane inlet Static pressure 2.72 Stage 2 deswirl vane exit Static pressure 2.90 Compressor exit ports Static temperature*, static pressure* * Used for performance evaluation Figure 4: 2.90 Compressor measurement locations. For the turbine, flow rate is measured with an orifice plate flow meter. Static temperature and static pressure are measured at the distribution manifold which supplies the turbine. Static pressures are also measured upstream of the turbine nozzle and in the diffuser section after expansion through the turbine. Static temperature and static pressure are finally measured at the exit of the rig. Table 2 lists the various measurements for turbine performance testing, and Figure 5 depicts a crosssection of the test rig which illustrates the measurement locations at the turbine end Copyright 202 by ASME

5 Table 2: Figure 5: Turbine measurement location descriptions. Station Physical Measurement No. Description Type 3.00 Turbine supply manifold Static temperature*, static pressure 3.0 Nozzle inlet Static pressure* 3.20 Nozzle exit Static pressure 3.30 Turbine exit Static pressure 3.4 Diffuser location (Not used for present analyses) 3.42 Diffuser location 2 Static pressure 3.43 Diffuser location 3 Static pressure 3.50 Turbine discharge ports Static temperature*, static pressure* 4.00 Orifice plate flow meter Mass flow rate (calculated)* 5.0 Turbine hub seal back pressure Static pressure * Used for performance evaluation Turbine measurement locations. One of the complications that was experienced was icing of the turbine discharge ports due to significant temperature drop of moist air while expanding through the turbine, especially at high-head, low-flow conditions of the compressor where power consumption was high. In addition to ice restricting the flow area, it also formed around the thermocouple probes, thus invalidating the static temperature measurements and turbine power calculations. However, turbine power could still be evaluated indirectly as the sum of measured compressor power and power losses from the system, i.e., bearing and windage losses. By operating the rig with a dummy disk replacing the compressor impellers, the turbine power requirement was low enough that the temperature drop due to expansion through the turbine did not cause icing, and the measured turbine power was equivalent to the bearing and windage losses at steadystate. Figure 6 shows the measured power loss data versus speed, which fits well with a cubic polynomial (R 2 = 0.995). The error bars on the data include one standard deviation of the measurement error and the instrumentation uncertainty error. It was assumed that the windage loss from the compressor wheels would be similar to that of the dummy disk. A validation point from a compressor test where turbine discharge temperatures were not compromised (i.e., causing icing at the turbine discharge), allowing direct evaluation of power loss as measured turbine power minus measured compressor power, is also plotted with the other data and shows good agreement with the model. Mean Power Loss (hp) Measured bearing/windage loss from turbine power with "dummy" disk replacing 0.7 compressor impellers Validation point: measured turbine power minus compressor power Speed (krpm) Figure 6: Measured power loss from bearing drag and shaft windage. 4 RESULTS AND DISCUSSION This section presents performance data (head, flow, and efficiency) for the SPARC compressor and turbine. Four speed lines at 84, 05, 0, and 5.5 krpm were recorded for both the compressor and turbine, although it is important to note that efficiencies at the 5.5 krpm speed line may be lower than actual efficiencies because the temperatures may not have reached steady state conditions (due to high synchronous vibration levels and limited air supply to sustain continuous operation at that speed, the rig was quickly operated at multiple points with only -2 minutes of operating time at each point). Each performance point is computed with a 0-second average of the raw data (sampled at khz) that were acquired after temperatures had reached a steady state at each operating point. The performance quantities were calculated using the methodology recommended in ASME PTC-0 [4]. Gas properties required for performance calculations (density, enthalpy, entropy, specific heat, etc.) were calculated using a moist air model [5][6] to account for the relative humidity in the ambient air (compressor inlet) and shop air supply (turbine inlet). For clarity, an overview of performance calculations is provided in this paper. 4. Performance Calculation Overview First, the flow rates for the compressor and turbine were calculated using venturi and orifice flow meters and the flow rate equation in [7]: Mean Power Loss (kw) m = C d E D 2 ρ ΔΔ t () β 4 The dummy disk was designed to have a near-identical overhung moment (i.e., weight times distance from center of gravity to inboard bearing centerline) compared to the compressor impellers in order to conserve the rotordynamic behavior. In Eq. (), C d is the orifice or venturi discharge coefficient (obtained as a function of Reynolds number from calibration), E is the expansion factor, D t is the throat or bore diameter of 5 Copyright 202 by ASME

6 the meter, ρ is the upstream fluid density, ΔΔ is the differential pressure at the meter, and β is the ratio of bore/throat diameter to upstream pipe diameter. Once the mass flow rate is known, the volumetric flow rate at the compressor and turbine inlet is calculated from: Q = m ρ ii (2) Here, ρ ii is the density at the compressor/turbine inlet. In order to compare with the predicted performance maps, it was necessary to convert measured temperatures and pressures into their total (stagnation) quantities at the compressor/turbine inlet locations and their static quantities at the compressor/turbine exit locations. The relationship between total and static pressures and temperatures is shown below: P t = P s + ρ s v 2 2 v2 (3) T t = T s + (4) 2 C p In Eqs. (3) and (4), the subscripts t and s refer to total and static quantities, respectively. The quantities ρ s, v, and C p are the static density, average fluid velocity, and specific heat of the moist air at the measurement location. Measured pressures are either total or static as described in Table and Table 2, but measured temperatures (T m ) are typically in between total and static quantities depending on the sensor s ability to recover the gas stream s kinetic energy. Thus, a recovery factor is defined as: r = T m T s T t T s (5) The recovery factor for a particular probe type depends on Mach number and must be obtained from calibration data. For these performance calculations, the recovery factor for an exposed junction thermocouple was estimated from the data in []. For pressures, the conversion between total and static quantities is straightforward if static temperature and pressure are measured since the static density can be easily obtained. If total pressures are measured and/or if gas velocities are high, iteration is required to calculate total and static quantities (readers are directed to [4] for more details). Once total and static pressures have been calculated at turbine and compressor inlet/exit locations, the actual head for the compressor is obtained from the difference in inlet total enthalpy and exit static enthalpy: a H cccc = h eeee h ii (6) As discussed previously, the actual head for the turbine was not calculated using this method since reliable temperature measurements could not be obtained at all test points (except for the validation case). Instead, the turbine actual head was calculated from the measured compressor data and bearing loss data with a dummy disk: a H tttt = m a cccc H cccc m tttt + Power Loss The polytropic head for the compressor is calculated from: n P P H cccc = np n P P eeee n P f P ii (8) P ii ρ ii Here, n P is the polytropic exponent and f is the Schultz polytropic head correction factor [9]. Once actual and polytropic head are known, the polytropic efficiency can be obtained from: P η cccc (7) = H P cccc (9) Ha cccc Finally, the adiabatic head for the turbine is calculated from: H tttt = h eeee h ii (0) Here, h eeee is the enthalpy of moist air at the exit pressure and inlet entropy. The adiabatic efficiency is equal to: η tttt = H tttt a () H tttt An additional analysis was performed on the measured data to quantify the uncertainty in performance due to temporal and spatial data scatter and instrument calibration uncertainty. The uncertainty analysis was performed by the basic Monte Carlo method [8] with 00,000 samples, and the standard deviation was calculated for each performance quantity at each test point. The basic Monte Carlo method uses random sampling of the normal distributions of the input variables (measurement data) to generate a distribution of output variables (performance). Table 3 lists the average standard deviation for each of the calculated performance quantities for the compressor and turbine measurements. Generally, the turbine performance data has more scatter than the compressor performance data. 6 Copyright 202 by ASME

7 Table 3: Average standard deviation of performance quantitites based on Monte Carlo analysis with 00,000 samples. Flow Head Efficiency Compressor 0.23% 6%.44% Turbine 0.96%.0% 2.4% 4.2 Compressor Performance The calculated compressor performance data are plotted next to the predicted performance maps provided by the compressor designer in Figure 7, Figure 8, Figure 9, and Figure 0. The prediction curves were based on a mean-line analysis method confirmed with computational fluid dynamics. Figure 7 and Figure 8 present the data in terms of head, flow, and efficiency, while Figure 9 and Figure 0 present the data in terms of head coefficient and flow coefficient. The data have been normalized to the predicted performance at the peak efficiency point at the design speed (05 krpm). These plots show reasonable correlation with the predicted curves, although the head is somewhat lower than predicted (approximately 0%). Efficiency is close to predicted values at 84 krpm, but it drops at higher speeds when the compressor discharge temperature rises above the boiling point of water, increasing the actual head produced by the compressor by the heat of vaporization of water. Data plotted in terms of head and flow coefficients (Figure 9 and Figure 0) nearly collapse on each other following the fan law. In the absence of Mach number and Reynolds number effects, disk friction, leakage, etc., the predicted curves would be expected to fall exactly on each other. Due to its small size, the relative assembly stack-up tolerances for clearances and alignment of the SPARC machine are generally higher than would be encountered on larger-scale equipment. For example, radial seal clearances were increased from 3 mil (0.076 mm) to 6 mil (2 mm) to alleviate the occurrence of rubbing during tests, and as a result, relative leakage through the seals would be higher than for a largerscale machine. Normalized Total-to-Static Polytropic Head Normalized Volumetric Flow 05.0 krpm, Predicted 0.0 krpm, Measured 5.5 krpm, Measured Figure 7: Polytropic head vs. volumetric flow for SPARC compressor. Normalized Total-to-Static Polytropic Efficiency Figure 8: Polytropic efficiency vs. volumetric flow for SPARC compressor. Normalized Total-to-Static Polytropic Head Coeff Normalized Volumetric Flow Figure 9: Polytropic head coefficient vs. volumetric flow coefficient for SPARC compressor. Normalized Total-to-Static Polytropic Efficiency Normalized Volumetric Flow Coeff Normalized Volumetric Flow Coeff. Figure 0: Polytropic efficiency vs. volumetric flow coefficient for SPARC compressor. 4.3 Turbine Performance 05.0 krpm, Predicted 0.0 krpm, Measured 5.5 krpm, Measured 05.0 krpm, Predicted 0.0 krpm, Measured 5.5 krpm, Measured 05.0 krpm, Predicted 0.0 krpm, Measured 5.5 krpm, Measured The calculated performance data are plotted next to the turbine designers predicted performance maps in Figure, Figure 2, Figure 3, and Figure 4. Again, the prediction curves were based on a mean-line analysis method confirmed with computational fluid dynamics. Figure and Figure 2 present the data in terms of head, flow, and efficiency, while Figure 3 and Figure 4 present the data in terms of head coefficient and flow coefficient. Again, the data have been 7 Copyright 202 by ASME

8 normalized to the predicted performance at the peak efficiency point at the design speed (05 krpm). Due to the surge and stonewall limits of the compressor, it was only possible to operate the turbine over a part of its performance map. As was the case for the compressor, the measured performance data trends agree reasonably well with predictions, although noticeable discrepancies exist. The turbine flow is higher than predicted, and the adiabatic efficiencies calculated are approximately 20-30% lower than predicted. The larger flow data compared to predictions is attributed to slightly larger flow area in the actual turbine geometry (approximately 0% larger). The change in flow would increase almost proportionally with area over a small range, so the normalized volumetric flow of the data is approximately 0% larger. It is well-known from experience that obtaining efficiency data from open-loop testing is more difficult than with closed-loop testing, partly due to large temperature gradients from one end of the machine to the other. For example, for some of the higher-speed, high compressor head test conditions, turbine exit temperatures were on the order of 0 F (-2 C) or lower while the compressor discharge exceeded 350 F (77 C). Physically, those two locations on the test rig are approximately 4-5 in. (00-25 mm) apart. In general, this points to questions about the confidence in the turbine data, also considering the icing issue in the turbine discharge discussed previously. Future testing with a closed-loop scheme is planned to obtain more reliable turbine data. As with the compressor predictions, the normalized coefficient curves for the turbine performance predictions (Figure 3 and Figure 4) do not collapse on each other exactly like the fan law. There is larger separation, especially at higher flow, and the test data also reflects this trend. Normalized Total-to-Static Adiabatic Head Normalized Volumetric Flow Figure : turbine krpm, Predicted 0.0 krpm, Measured 5.5 krpm, Measured Adiabatic head vs. volumetric flow for SPARC Normalized Total-to-Static Adiabatic Efficiency Figure 2: Adiabatic efficiency vs. volumetric flow for SPARC turbine. Normalized Total-to-Static Adiabatic Head Coeff Normalized Volumetric Flow Figure 3: Adiabatic head coefficient vs. volumetric flow coefficient for SPARC turbine. Normalized Total-to-Static Adiabatic Efficiency Normalized Volumetric Flow Coeff Normalized Volumetric Flow Coeff. Figure 4: Adiabatic efficiency vs. volumetric flow coefficient for SPARC turbine. 5 CONCLUSIONS 05.0 krpm, Predicted 0.0 krpm, Measured 5.5 krpm, Measured 05.0 krpm, Predicted 0.0 krpm, Measured 5.5 krpm, Measured 05.0 krpm, Predicted 0.0 krpm, Measured 5.5 krpm, Measured Some high GOR wells are currently not competitive to produce with conventional equipment. A Subsurface Process And Reinjection Compressor (SPARC) tool is under development which will process, compress, and re-inject up to 30% of the gas that comes out of these formations, resulting in higher maintained reservoir pressures and increased well production at a fraction of the cost of a surface-based operation. 8 Copyright 202 by ASME

9 An aerodynamic performance test stand was developed for validation of the performance of the turbomachinery a twostage centrifugal compressor driven by a single-stage mixedflow expansion turbine. The test stand is configured for openloop testing of the overall compressor and turbine performance with air as the working fluid and atmospheric pressure at the compressor suction and turbine discharge locations. The test rig was instrumented as extensively as possible, given the relatively small size of the device, with measurement redundancy employed at critical performance measurement locations. Gas properties were calculated using a moist air model to account for the relative humidity in the ambient air (compressor inlet) and shop air supply (turbine inlet). Several performance curves were generated for each component ranging from krpm (70-0% of design speed). In general, measured compressor head was approximately 0% lower than predictions; measured efficiencies for the 84 krpm speed line were very close to the predicted values. The turbine had higher flow than predictions, due in part to a larger flow area in the manufactured turbine wheel. It was discussed that efficiency measurements obtained from open-loop test rigs are often difficult due to large temperature gradients of the components, the present case included. In all, the SPARC tool has shown acceptable performance on the open-loop test stand. It is possible that the design speed would need to increase approximately 5% to make up for the difference in compressor head, but it was demonstrated that this should be within the machine s capabilities. Future work entails higher-pressure, closed-loop testing with nitrogen as the working fluid and integrated testing with the complete SPARC tool (i.e., separators, bypass valve, recycle valve, etc.) followed by implementation of tool at a down-hole test site. The higher-pressure testing will be used to verify aerodynamic performance data and will provide more accurate data, especially for the turbine, in accordance with an ASME PTC-0 Type 2 test. These activities will be the subject of future publications. 6 ACKNOWLEDGMENTS The compressor and turbine aerodynamic design and performance predictions were provided by Turbo Solutions Engineering. The authors wish to thank the Prudhoe Bay Unit owners BP, Exxon Mobil, and ConocoPhillips; Halliburton; and Southwest Research Institute for the opportunity to publish this work. While this paper reflects the views of its authors, it may not necessarily reflect the views of BP and/or Exxon Mobil and/or ConocoPhillips and/or Halliburton and/or Southwest Research Institute and/or Turbo Solutions Engineering. 7 NOMENCLATURE The following is a listing of the nomenclature presented in this paper. Units are presented in terms of mass [M], length [L], time [T], and temperature [K]. Arabic C d Orifice or venturi discharge coefficient [-] C p Specific heat at constant pressure [L 2 T -2 K - ] E Expansion factor [-] f Schultz polytropic head correction factor [-] H Head [ML 2 T -2 ] h Enthalpy [ML 2 T -2 ] m Mass flow rate [MT - ] n P Polytropic exponent [-] P Pressure [ML - T -2 ] ΔP Pressure differential [ML - T -2 ] Q Volumetric flow rate [L 3 T - ] r Thermocouple kinetic energy recovery factor [-] T Greek Temperature [K] v Velocity [LT - ] β Subscript Orifice or venturi bore diameter-to-upstream diameter ratio [-] η Efficiency [-] ρ Density [ML -3 ] s t cccc tttt Superscript a P Static Total or stagnation Compressor Turbine Actual Polytropic Adiabatic 8 REFERENCES [] Brady, J.L., Klein, J.M., Stevenson, M.D., Petullo, S.P., D Orsi, N.C., Skinner, R.C., Eager, K.D., and Perry, R.A., 998, Downhole Gas Separation and Injection Powered by a Downhole Turbo Expander, Proceedings of SPE Annual Technical Conference and Exhibition, SPE 4905, New Orleans, Louisiana, USA, September Copyright 202 by ASME

10 [2] DellaCorte, C., Valco, M.J., 2000, Load Capacity Estimation of Foil Air Journal Bearings for Oil-Free Turbomachinery Applications, Tribology Transactions, Vol. 43, pp [3] Picaro, A., Childs, D., 2005, Rotordynamic Coefficients for a Tooth-on-Stator Labyrinth Seal at 70 Bar Supply Pressures: Measurements Versus Theory and Comparisons to a Hole-Pattern Stator Seal, ASME Journal of Engineering for Gas Turbines and Power, 27, pp [4] ASME, 997, Performance Test Code on Compressors and Exhausters, PTC-0. [5] Nelson, H.F., Sauer H.J., 2002, Formulation of High- Temperature Properties for Moist Air, HVAC&R Research, Vol. 8 (3), pp [6] Conde, M., 2007, Thermophysical Properties of Humid Air: Models and Background, M. Conde Engineering, Zurich. [7] ASME, 995, "Measurement of Fluid Flow in Pipes Using Orifice, Nozzle, and Venturi," MFC-3M-989. [8] Benedict, R.P., 959, Temperature Measurement in Moving Fluids, ASME Paper No. 59-A-257. [9] GMRC, 2006, "Guideline for Field Testing of Gas Turbine and Centrifugal Compressor Performance," Release 2.0. [0] Haldar, A., Mahadevan, S., 2000, Probability, Reliability, and Statistical Methods in Engineering Design, Wiley, New York. 0 Copyright 202 by ASME

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