Two-phase operation of a Terry steam turbine using air and water mixtures as working fluids

Two-phase operation of a Terry steam turbine using air and water mixtures as working fluids

Journal Pre-proofs Two-phase Operation of a Terry Steam Turbine using Air and Water Mixtures as Working Fluids Abhay Patil, Yintao Wang, Matthew Solom...

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Journal Pre-proofs Two-phase Operation of a Terry Steam Turbine using Air and Water Mixtures as Working Fluids Abhay Patil, Yintao Wang, Matthew Solom, Ashraf Alfandi, Shyam Sundar, Karen Vierow Kirkland, Gerald Morrison PII: DOI: Reference:

S1359-4311(19)32577-3 ATE 114567

To appear in:

Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

9 April 2019 10 September 2019 19 October 2019

Please cite this article as: A. Patil, Y. Wang, M. Solom, A. Alfandi, S. Sundar, K. Vierow Kirkland, G. Morrison, Two-phase Operation of a Terry Steam Turbine using Air and Water Mixtures as Working Fluids, Applied Thermal Engineering (2019), doi:

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© 2019 Published by Elsevier Ltd.

Two-phase Operation of a Terry Steam Turbine using Air and Water Mixtures as Working Fluids

Abhay Patil1, Yintao Wang1, Matthew Solom2, Ashraf Alfandi1, Shyam Sundar1, Karen Vierow Kirkland1, Gerald Morrison1


A&M University, 2Sandia National Laboratories*

Highlights 1. A Terry ZS-1's performance was tested for gas mass fractions from 1.0 to 0.05. 2. Terry turbines show a strong adaptability under two-phase flow operating conditions. 3. The expansion of two-phase mixtures changes from isentropic to near isothermal as the air mass fraction varies from 1 to 0.05. 4. The turbine performance systematically degrades with increasing liquid content.

Sandia National Laboratories is a multi-mission laboratory managed and operated by National Technology & Engineering Solutions of Sandia, LLC, a wholly owned subsidiary of Honeywell International Inc., for the U.S. Department of Energy’s National Nuclear Security Administration under contract DE-NA0003525. This paper describes objective technical results and analysis. Any subjective views or opinions that might be expressed in the paper do not necessarily represent the views of the U.S. Department of Energy or the United States Government. SAND2019-XXXX *


Abstract Terry steam turbines are employed in the safety systems of many nuclear Boiling Water Reactors to drive pumps and provide cooling water to the nuclear reactor core. While the turbine efficiency is low, the more important feature is high reliability under off-normal conditions. An important aspect of reliability is the ability to function with two-phase steam-water injection into the turbine, as most likely occurred in the Fukushima Dai-ichi nuclear accidents. This study investigates the characteristics of a Terry turbine during air-water injection with gas mass fractions ranging from 1 (dry gas) to 0.05 (wet gas), to better understand the Terry turbine’s true operational capabilities and provide justification for extended Terry turbine use for reactor safety. Other parameters investigated are the inlet pressure, the exhaust backpressure and the turbine’s rotational speed. The turbine performance is presented in terms of dynamometer loading and pump performance change as functions of the gas mass fraction.




Terry Turbines and the RCIC System The RCIC system is an engineered safety feature that provides makeup water for core cooling

of many Boiling Water Reactors (BWRs) with a Mark I or later containment. The RCIC system consists of a steam-driven Terry turbine that powers a pump to provide makeup water to the reactor pressure vessel (RPV). The turbine takes steam from one of the Main Steam Lines exiting from the RPV and exhausts to the Suppression Pool in the containment. The RCIC pump takes suction from the Condensate Storage Tank or Suppression Pool and delivers this water to the RPV for core cooling.


Post-Fukushima Dai-ichi accident investigations found that the extended operation of the RCIC system played a major role in mitigating the accident consequences in Unit 2 by removing core decay heat for about 70 hours [1]. Originally, the RCIC system was designed to shut down at a specific RPV high water level to prevent reactor overfill and the resulting introduction of liquid water to the Main Steam Line. Such introduction leads to water ingestion in the turbine, which lowers turbine as well as system performance efficiency. Due to the Station Blackout conditions and loss of DC power, the RCIC system controller was unable to regulate the governor valve position, which resulted in an uncontrolled operation out to 70 hours. The system eventually shut down, with a postulated cause being overheated turbine oil and consequent turbine bearing failure resulting from high Suppression Pool temperatures [2].


The Need for Data on Two-phase Operation of the Terry Turbine The RCIC System ran for most of these 70 hours without turbine speed regulation by the

governor valve. The water level in the RPV is hypothesized to have risen enough to spill over into the Main Steam Line, resulting in steam-water two-phase flow into the Terry turbine. While the Terry turbine has been shown to be successful in withstanding water slug injection [3], less is known about its ability to recover and enable sustained operation of the RCIC System after the water slug passes through. Characterization of the Terry turbine performance under two-phase conditions is necessary to understand the potential of the RCIC System to remove decay heat under accident conditions when regulation of the turbine speed is not available, such as Station Blackout and Extended Loss of AC Power (ELAP) scenarios. The database presented here is an improvement over past databases. As reviewed in the next section, some air-water and steam-water data have been published. The data acquired for the


current data base are important because testing was conducted in finer increments over the parameter ranges, allowing for greater accuracy in performance characterization of the Terry turbine. Another value of this study is that the data acquired herein will be used to develop analytical models of the Terry turbine with one application being for reactor safety analyses. The data will also be utilized to scale from the smaller scale Z-series Terry turbines to the full-scale G-series turbines used in nuclear power plants and operating with steam as the working fluid. Specifically, the data herein may be used to validate and further improve the new scaling methodologies that have been developed for the RCIC System’s turbomachinery [4, 5]. The data obtained herein will also be useful for Pressurized Reactor Water (PWR) safety analyses, as PWRs use Terry turbines for their Turbine-driven Auxiliary Feedwater System which provides water to the steam generators on the secondary coolant loop.


Previous Knowledge Methodologies for the performance characterization of turbomachines using single phase

fluids are well established and follow the standard affinity between the output parameters. Serrano et. al. [6] proposed a new methodology to calculate characteristic curves of twin entry turbines working under different inlet conditions. Usually, efficiency is defined based on a pure isentropic expansion process. Ba et al. [7] modified and proposed a new definition of efficiency for cooled turbines based on a mixed expansion process. Weiß et al [8] experimentally investigated axial impulse turbine and a radial cantilever turbine to establish a performance benchmark for fluctuating heat fluxes and partial load conditions. Tahani et al [9] developed a wet compression thermodynamic model that considers cooling down the compressor inlet air to increase turbine


efficiency. Input data such as droplet diameter, amount of overspray and droplet size were used to match the compressor work with the turbine output. Trela et al [10] utilized exergy efficiency to characterize the losses associated with two-phase interactions in a steam nozzle. Öhman and Lundqvist [11] performed an experimental investigation of a Lysholm Turbine operating with superheated, saturated and 2-phase inlet conditions. Interestingly, they found that the vapor quality does not affect the turbine power output. Date et al [12, 13] performed two-phase experiments of a reaction turbine to understand the effects of feedwater temperature and nozzle configurations. An increase in feedwater temperature improved the isentropic efficiency. Yang et al [14] proposed a new theoretical synchronal rotary multiphase pump (SRMP) model to predict pump performance at very high inlet gas volume fractions (GVFs). The model, supported by experimental investigation, shows the operational capability of the SRMP at very high inlet GVFs. While there are studies focused on two-phase interactions of different conversion devices, published testing results from Terry turbines are relatively limited, especially for two-phase flow conditions. Before the Fukushima accident, the performance map and operational limits of the turbine were mostly based on information provided by manufacturers. Assumptions and system models in reactor safety codes were based on the rated performance of the RCIC System’s Terry turbine and pump [15], [16]. The RCIC system performance under Beyond Design-Basis Event (BDBE) conditions continues to be largely unknown and based on conservative assumptions used in Probabilistic Risk Assessment (PRA) applications. This prompted a series of research studies to explore the RCIC system’s performance under BDBE conditions. The development of mechanistic models and scaled-down testing of RCIC components under multiphase flow were the focus of preliminary studies. The likely state and accident progression of Unit 3 six days into the accident was studied by Fernandez-Moguel and Birchley [17] using the


reactor safety code MELCOR 2.1. Li et al. [18] performed a sensitivity analysis of the components modeled for the RCIC system, thermal stratification of the suppression chamber, and mixing flow and sea water injections in Unit 2. Lopez et al. [19, 20] used RELAP/ScdapSIM to model the Unit 2 accident and studied the effects of RCIC system operations on the accident progression. The degradation of turbine performance was evaluated using a degradation coefficient, representing turbine power reduction as a function of the steam flow quality. The self-regulating mode of the RCIC system and the feasibility of extending its operating range was investigated by Zhao et al. [21]. The models employed by Zhao et al. [22] are able to predict the steam flow rate and supersonic velocity at the Terry turbine bucket inlet; these input models of Terry turbine-driven RCIC systems predict their behavior under normal operating conditions adequately. However, these models are not able to represent Terry turbine behavior with confidence under two-phase conditions. An aim of the current study is to provide much-needed physical test data for model development. Luthman [23] developed the turbine test rig and performed limited initial investigations for determining the performance curves of a Terry turbine with two-phase inlet flow. Both air-water and steam-water data were presented for a ZS-1 Terry turbine. For the steam tests, data were presented for several inlet flow qualities. Torque output using steam was higher compared to air as a working medium, although a significantly wider range of data for two-phase flow was needed to be conclusive. The data presented in this paper were acquired using the same equipment as Luthman [23]. Lutman’s techniques were modified to include a wide range of two-phase conditions that cover a broader set of hypothesized operating conditions as well as to provide clearer identification of the torque, power and efficiency trends. Testing in finer increments over


the parameter ranges allows for greater accuracy in the performance characterization of the Terry turbine.


The Terry Turbine The Terry steam turbine used in the RCIC System is an impulse turbine, similar to the Pelton

wheel by design. In the ideal case of dry steam, the turbine would follow the characteristic curve (provided by manufacturer) for any change in the inlet conditions. When water is present, the expansion process across the nozzle can be significantly different from the case of dry steam and can degrade the turbine performance. The workings of a Terry turbine are shown in Figure 1. Steam discharged through the converging-diverging nozzle impinges on the buckets in the turbine wheel, imparting kinetic energy to the wheel. The steam exiting a bucket then enters “reversing chambers” mounted on the inner surface of the turbine casing. The steam reverses direction in the reversing chambers and is redirected back into the buckets to deposit more energy onto the wheel. The resulting spiral flow of the steam is depicted in the left side of Figure 1.


Figure 1: Terry turbine (left) and rotor and nozzle (right) [3] 8

The precise design parameters of many Terry turbines in industrial use are unavailable because their lineages are unclear. The turbine used for testing here, as is common, has no plate or other identification. However, much is known about its capabilities from experience and extensive consultation with the steam turbine experts at Revak Keene Turbomachinery in La Porte, TX. A Terry turbine model ZS-1 was used in the current work. The Z-series Terry turbines (i.e., models ZS and Z) have a wheel diameter of 18 inches, while G-series (models GS and G) have a 24-inch wheel. US nuclear power plants employ either a GS-1N or a GS-2N model Terry turbine in the RCIC System. The GS-series turbines differ in their maximum number of steam inlet nozzles. The ZS-1 Terry turbine utilized in the current experiments has ports for up to four steam inlet nozzles. One nozzle is installed and the other three ports are plugged to establish the baseline performance. Figure 2 shows the turbine lower casing and, separately, the assembled test skid installed in the test loop. The nozzle has a nominal throat diameter of 0.380 inch. The trip speed is 4,100 rpm and its standard turbine rotational rate is 3,600 rpm. The governor system has been removed. Since the turbine utilizes only one nozzle, this configuration is desirable for the establishment of performance benchmarks for scaling laws to represent the degradation as a function of two-phase flow at the lower end of the size and capacity scale for Terry turbines.


Figure 2: ZS-1 Terry turbine a) Nozzle location [23] b) Current set-up for two-phase testing

2. Experimental Test Setup 2.1

Experimental Facility Figure 3 shows the schematic of the multiphase flow loop used to test the turbine. The flow

loop consists of water and air lines. The air supply comes from oil-free screw compressors with a total capacity of 4250 m3/hr at 0.83 MPa. Air filters and dryers are installed in-line to maintain the cleanliness and humidity of the supply air. Two air flow meter lines are used to measure and record different ranges of flow rates. A 5000-gallon water tank is filled with city water before the test. During the test, a centrifugal pump circulates filtered water through the test loop. A total of 4 lines are available to supply the required flow rate based on specific flow meter capacity. Not all the lines were used due to the capacity limits of the turbine under consideration. Due to the complexity in the flow loop and turbine plenum path, the flow homogenization and liquid droplet size are not controlled. The ideal gas law and isentropic expansion are utilized to characterize flow parameters.


The detailed operating processes including the control of inlet pressure, air mass fraction at the turbine inlet and rotational speed are explained in Section 3.


Figure 3: Schematic of experimental facility to test Terry turbine 12

Table 1: Equipment calibrations and certifications Sensor Identification and Range

Where Used in Facility


Calibration Standard

Pressure Transmitter Rosemount 0-2.07 MPa

Air inlet

±0.04% of span

ISO 10474 3.1 EN 10204 3.1 Cal system: Z540-1-1994

Pressure Transmitter Rosemount 0-1.03 MPa

Turbine inlet

±0.04% of span

ISO 10474 3.1 EN 10204 3.1 Cal system: Z540-1-1994

Pressure Transmitter Rosemount 0-0.517 MPa abs

Turbine exhaust chamber

±0.04% of span

ANSI Z540-1-1994 Cal system: Z540-1-1994

Pressure Transmitter Rosemount 0.517 MPa abs


±0.04% of span

ANSI Z540-1-1994 Cal system: Z540-1-1994

Pressure Transmitter Rosemount 0-0.69 MPa

Turbine outlet pipe

±0.04% of span

ISO/IEC 17025:2005

< 2.5 °C max, 1°C typical + thermocouple error (Type T near room temperature) 6.230 μV @ fullscale readings when scale is ±10 V

ANSI/NCSL Z540.1-1994 (R2002)

NI 9213 Module


NI 9205 Module


Load cell (Omega)


Linearity: ±0.03% FSO

NIST standard

Gas Turbine Flowmeter

Air supply

±1% of RD

ISO 17025:2005

Foxboro Vortex Flowmeter

Air supply

±1.0% of rate

ISO 17025:2005

Tachometer (Monarch ACT-3X) (5-500,000 RPM)


± 0.001% of reading


Badger Flow meter (M2000)

Water line

0.20% + 1 mm/s error

ISO/IEC 17025:2005

Coriolis flow meter

Water line

±0.10% of span

FSO = Full Scale Output RD = Reading 13

ANSI/NCSL Z540.1-1994 (R2002)

The sensor types and locations are given in Table 1. A Stuska XS-19 water brake dynamometer is used to load the turbine shaft. The dynamometer can operate at speeds up to 120,000 rpm with loading up to 149 kW at that speed. During a test, the load on the turbine is controlled by regulating the water supply to the dynamometer. An electrically controlled needle valve (CV-1) is used to precisely control the flow rate to the dynamometer. All the sensors were integrated on a National Instruments platform of data acquisition cards with LabVIEW 2015 used for PID control, data monitoring, and recording. The GUI (Graphical user interface) was designed to aid the user to monitor, record, and control the flow parameters.


Method of Uncertainty Evaluation The sensors employed were calibrated to the standards required of the project at external

calibration facilities. The uncertainty associated with the calculation of desired parameters was estimated based on the errors in the measurement of known parameters using the KlineMcClintock method [24]. For example, the uncertainty associated with the measurement of input power Pgas is given by Eq. 1, where p is a pressure and V is the volumetric flow rate

𝑈𝑃𝑔𝑎𝑠 =




) (

∂𝑝𝑖𝑛𝑙𝑒𝑡 𝑈𝑝𝑖𝑛𝑙𝑒𝑡



) (


∂𝑉𝑖𝑛𝑙𝑒𝑡 𝑈𝑉𝑖𝑛𝑙𝑒𝑡



2 1/2


∂𝑝𝑜𝑢𝑡𝑙𝑒𝑡 𝑈𝑝𝑜𝑢𝑡𝑙𝑒𝑡


Similarly, uncertainties associated with other parameters were calculated and values are shown in Table 2. Table 2: Uncertainty associated with desired parameters Outcomes Uncertainty Efficiency 2% Power Output 1% Input Gas Power 1.85% 14

Input Liquid Power 3.

Operating Procedures and Test Conduct


Single-phase Tests


Flow Control (Power input): For the single-phase tests, dry air was supplied at the turbine inlet. The supplied air’s relative humidity was about 20%. Air pressure was controlled using the electropneumatic control valve ACV-1. PID control was implemented to ensure constant turbine inlet pressure. The turbine was operated over the inlet pressure range of 0.21 MPa to 0.76 MPa. Flowmeters installed before ACV-1 measure the air flow rate at a constant pressure. Pressure and temperature sensors are employed at the turbine inlet and exhaust. The ideal gas law is utilized to calculate air density and volume flow rate at the turbine inlet. Shaft Power Control (power output): Dynamometer loading was used to control the turbine shaft speed. Water flow through the dynamometer regulates its loading; increased flow increases the amount of torque it absorbs. Flow control to the dynamometer is achieved by electrically controlled needle valve. The desired turbine speed is obtained by adjusting the needle valve; an increase in the load would slow the turbine’s rotational speed. During facility shakedown tests, the temperature change across the turbine was observed to be a function of air mass fraction. When dry air was supplied to the turbine (typically at or slightly above room temperature), the energy removed by the turbine from the air at modest inlet pressures resulted in modest temperature drops; the temperature drop from inlet to exhaust increased with inlet pressure. At high inlet pressures, the exhaust temperature dropped well below freezing in several cases. The MELCOR code does not behave well at or below freezing temperatures. Therefore, an air heater was employed to preheat the air to ensure the turbine outlet temperature does not fall below freezing. In the testing after shakedown, the data were recorded once a steady15

state condition was achieved. Specifically, the inlet pressure, inlet temperature, and rotational speed were held constant with time. Data were recorded at steady state for 30 to 45 seconds at 3 samples/second. Tests were terminated by turning off the air valve first and then the dynamometer needle valve once the shaft had come to rest.


Two-phase Tests For two-phase flow tests, the turbine was loaded by supplying water to the dynamometer. The

desired turbine inlet pressure was obtained by automatic adjustments of the air electro-pneumatic control valve ACV-1. The booster pump was started to supply the water. The appropriate water valve was then opened based on the required flow rate to maintain the desired air mass fraction at the turbine inlet. The dynamometer load was adjusted to obtain the desired rotational speed. Data were recorded in a similar fashion once the desired conditions were achieved. Tests were concluded by first closing the air and water valves to the turbine inlet. The dynamometer needle valve was then closed once the shaft came to rest.


Test Ranges

The two-phase performance of the employed Terry turbine was explored for the conditions given in Table 3. Table 3: Operating Conditions Inlet Pressure (MPa)

Air mass fraction

Rotational speed (RPM)




As noted in Section 3.1, the inlet temperature was maintained at 305 K to avoid freezing temperatures at the turbine outlet. For two-phase flow conditions, the increased liquid content 16

had a greater influence on the inlet temperature as well as the two-phase flow expansion process. Figure 4 shows the inlet temperature values for different air mass fraction conditions.

Figure 4: Turbine inlet temperature at different inlet pressures and air mass fractions


Performance Mapping and Data Analysis The results include the overall performance map for single-phase and two-phase flow tests. To

characterize the turbine performance, torque and efficiency curves are generated. An energy balance is utilized to evaluate the work performed by the Terry turbine. The power input has two components for two-phase flow: the energy imparted by the gas, and the energy imparted by the liquid. The gas input power is modeled using an isentropic expansion (Equation 4) while the liquid input power is modeled by Equation (5). Here, ∆𝑝 is the pressure difference across the turbine, and 𝑄𝑙 is the volumetric flow rate of the liquid. Torque is the desired output 17

parameter which is calculated by using 𝛤 = 𝐹 ∙ 𝑟, where F is the loadcell output and r is the arm length of dynamometer. Isentropic efficiency is given by ƞ𝑖𝑠𝑒𝑛𝑡𝑟𝑜𝑝𝑖𝑐 =

𝑃𝑜𝑢𝑡𝑝𝑢𝑡 𝑃𝐼𝑛𝑝𝑢𝑡



= 𝑃𝑔𝑎𝑠 + 𝑃𝑙𝑖𝑞𝑢𝑖𝑑

where Turbine Power Output: Gas Power Input: Liquid Power Input:



𝑊𝑠ℎ𝑎𝑓𝑡 = 𝛤 ∙ 𝜔 𝑘―1



𝑃𝑔𝑎𝑠 = (1 ― 𝑘)𝑝𝑖𝑛𝑙𝑒𝑡𝑉𝑖𝑛𝑙𝑒𝑡


𝑝𝑜𝑢𝑡𝑙𝑒𝑡 𝑝𝑖𝑛𝑙𝑒𝑡




𝑃𝑙𝑖𝑞𝑢𝑖𝑑 = 𝑄𝑙 ∙ ∆𝑝

(4) (5)

Single-phase (Dry Air) Performance Maps and Analysis Figure 5 shows the performance parameters, specifically torque and efficiency, at an air mass

fraction of 1.0. Increased power input from increased flow rates and pressure differences across the turbine correlates to increased torque and efficiency. Starting from the highest torque value at the lowest shaft rotation rate, the torque decreases linearly as the speed increases per the characteristic performance curve. The mechanical power extraction, torque times speed, is zero when either the torque or the shaft speed is zero. For a given inlet pressure, the optimum power output is achieved when the imparting jet produces a maximum amount of work with a minimum amount of incurred flow losses. Flow losses increase at non-optimum conditions such as changes in turbine load, fluid properties (a two-phase condition in this case) and operational conditions (inlet pressures and temperatures). Due to this the efficiency typically increases first, achieves a maximum value, and decreases with further increases in speed. With an increase in inlet pressure, the optimum output power increases with further increases the rotational speed.


There are notable efficiency jumps when the inlet pressure is increased from 0.21 to 0.345 MPa and from 0.345 to 0.48 MPa that are not observed when the inlet pressure is increased above 0.48 MPa in similar increments. This is partially attributed to the frictional drag of the turbine, which is primarily a function of turbine speed and not of inlet pressure. Hence as the energy supplied by the air increases with increasing pressure a smaller fraction of the air’s energy is consumed overcoming the parasitic friction. Another reason is the change in the expansion process in the nozzle changing the fluid momentum; low pressure differentials across the nozzle allow for the formation of a detrimental shock front or even purely subsonic flow that reduces the turbine efficiency.

Figure 5: Torque and efficiency vs rotational speed for tests with air mass fraction of. 1.0 4.2

Two-phase Performance Maps and Analysis


Gas Volume Fraction vs Gas Mass Fraction through a single nozzle

Before proceeding to the two-phase analysis, it is important to note the quantitative difference between the volume fraction of the gas versus the mass fraction of the gas. Equations (6) and (7) define the Gas Volume Fraction and gas Mass Fraction, respectively. 𝑉𝑔

𝐺𝑎𝑠 𝑉𝑜𝑙𝑢𝑚𝑒 𝐹𝑟𝑎𝑐𝑡𝑖𝑜𝑛(𝐺𝑉𝐹) = 𝑉𝑔 + 𝑉𝑙 19



𝐺𝑎𝑠 𝑀𝑎𝑠𝑠 𝐹𝑟𝑎𝑐𝑡𝑖𝑜𝑛(𝑀𝐹) = 𝑀𝑔 + 𝑀𝑙


The large difference in fluid density between air and water causes a multi-order relationship between GVF and MF. A Mass Fraction of 0.005 is equal to 0.52 GVF (presenting an unstable or slug flow regime) while a Mass Fraction of 0.05 and more refers to a GVF of more than 0.9. Figure 6 shows the relationship between MF and GVF for air and water at 0.48 MPa and room temperature.

Figure 6: Gas volume fraction vs gas mass fraction at 0.48 MPa 4.2.2 Sonic speed for two-phase flow At a nozzle throat, the speed of a compressible flow cannot exceed the local speed of sound. If there is a sufficient pressure drop across the nozzle, a Mach number greater than 1 may be present at the outlet. The sonic speed of a two-phase mixture is greatly affected by the flow homogeneity, extent of thermodynamic equilibrium, and pressure pulses. For the reader’s reference, a discussion of the characteristics of the sonic speed of homogeneous two-phase flow is presented before proceeding to the experimental data. A general


understanding of two-phase sonic speed and its effect on the momentum of a jet will help provide insight into the two-phase flow performance of a Terry turbine. The sonic speed of a two-phase liquid/gas mixture can be expressed by Equations (8) and (9) [25], in which GVF is the gas volume fraction: 1 2


= [𝜌𝑙(1 ― 𝐺𝑉𝐹) + 𝜌𝑔𝐺𝑉𝐹]



𝐺𝑉𝐹 (1 ― 𝐺𝑉𝐹) + 𝑘𝑝 𝜌𝑙𝑐2𝑙


where c is the sonic speed, k is the isentropic index and ρ is the fluid density.

In many applications,

𝐺𝑉𝐹 𝑘𝑝

(1 ― 𝐺𝑉𝐹) 𝜌𝑙𝑐2𝑙


= 2


thus Equation 8 can be simplified to 𝐺𝑉𝐹 ∙ [𝜌𝑙(1 ― 𝐺𝑉𝐹) + 𝜌𝑔𝐺𝑉𝐹] 𝑘𝑝


Figure 7 shows the sonic speed for two-phase air-water flow as a function of different GVF at the pressures used in the current set of turbine tests. The sonic speed of a homogenous two-phase flow first decrease and subsequently increases with increases in GVF. The decrease in sonic speed in the two-phase flow will weaken the jet momentum, effectively reducing the turbine power output since the volumetric flow rate is limited by the sonic speed and nozzle throat area.


Figure 7: Sonic speed of two-phase water-air mixtures at five relevant pressures


Variation in Power Output and Input as a function of two-phase flow

The changes in desired output parameters such as torque and power output are presented in Figure 8 as functions of air mass fraction and inlet pressure. The torque performance is seen to increasingly degrade with increased water content in the two-phase mixture. With decreases in inlet pressure, the torque values decrease due decreased power input. The increased presence of liquid has a large impact on the torque output with the performance envelope shrinking as the liquid fraction increases.


Figure 8: Effect of air mass fraction and inlet pressure on two-phase flow performance (torque) of the ZS-1 Terry turbine. Air mass fraction: a) 1.0, b) 0.70, c) 0.40 d) 0.05

The next task was to explore how a change in air mass fraction at a given inlet pressure affects the turbine input and output performance. Figure 9 a) shows the total power input as a function of the air mass fraction at an inlet pressure of 0.483 MPa. The increased presence of liquid droplets reduces the total flow rate and also affects the expansion of the mixture across the nozzle. With increases in liquid content, the power input decreases consistently. Increased liquid content also affects the jet momentum due to a reduction in the sonic speed as illustrated by Figure 7, limiting


the performance envelope of the turbine. Please note that the sonic speed in Figure 7 is a function of GVF while the turbine performance is characterized in terms of the air mass fraction. Starting from low rotational speeds, the power output increases and marches towards the optimum value as the rotational speed increases. The output achieves an optimum value at a specific rotational speed; this is the design point, or the best efficiency point, where the losses are at a minimum. After achieving this optimum value, the power output decreases with further increases in rotational speed. This is per the characteristic output curve of the turbomachines. With decreases in the air mass fraction, the power input decreases due to reduced mixture velocities, which further reduce the power output while additionally shifting the operating envelope and design point toward lower rotational speeds (Figure 9b). The lowest point of the output power at higher rotational speeds indicates the minimum power required to overcome frictional drag offered by bearings and seals and the minimum loading point of the dynamometer. The resulting minimum power output is around 0.15 kW for the current ZS-1 test skid.


Figure 9: a) Power input; b) Power output at 0.483 MPa inlet pressure for various air mass fraction values Figure 10 a) shows the power input as a function of air mass fraction at 0.76 MPa. With an increase in pressure, the total energy input is increased. Increased inlet pressure also helps in homogenizing the flow, effectively reducing the momentum losses and increasing the sonic speed 25

of the two-phase mixture. The overall effect is an increase in the performance envelope as highlighted by the increased power at all corresponding points. This is evidenced by an apparent reduction in triangle size in Figure 10 a) indicating improved nozzle performance. Increased inlet pressure also results in an increased power output (Figure 10 b). Furthermore, increased inlet pressure moves the best efficiency point (BEP) towards a higher rotational speed. The minimum dynamometer loading represents the bearing and friction drag. It is a function of rotational speed and its value is around 0.17 kW at 3000 rpm. A triangle presenting the minimum dynamometer loading apparently looks smaller than in Figure 9 b) due to the improved performance map at the higher inlet pressure.


Figure 10: a) Power input, b) Power output for an inlet pressure of 0.76 MPa

Figure 11 presents the thermodynamic efficiency as a function of air mass fraction. The overall efficiency decreases with a decrease in the air mass fraction. The best efficiency point moves 27

towards lower speeds for lower air mass fractions due to reductions in jet momentum; the shifting BEP positions for high air content tests is illustrated by the red arrow on the figure. The performance map consistently shrinks (indicated by the blue arrow) and the BEP moves to lower rotational speeds.

Figure 11: Efficiency at 0.483 MPa inlet pressure for different air mass fractions

The BEP decreases consistently with decreasing air mass fractions until the air mass fraction falls below about 0.10. At that point, the efficiency begins to recover. This rise increases with increases in inlet pressure (Figure 12). Closer inspection reveals a consistent decrease in the power output values at these conditions with the BEP still moving towards the left and a narrowing of the operating envelope, indicating that this apparent increase may be due to the applied definition of efficiency. It seems the expansion process transitions from isentropic to isothermal as the liquid presence increases, which may require a modified equation for efficiency evaluations at lower air


mass fractions. As the water content increases, the thermal capacitance also increases, and the water content provides heat to maintain the air temperature.

Figure 12: Isentropic efficiency at 0.76 MPa for varying air mass fractions

Figure 13 presents the relative change in temperature with respect to inlet temperature as air mass fraction varies. Larger temperature changes indicate greater expansion across the nozzle. The temperature change increases with increases in inlet pressure.

The temperature change is

insignificant at all tested inlet pressures for air mass fractions lower than 0.40. This indicates a polytropic expansion process for the air, which is nearly isothermal at lower air mass fraction values.


Figure 13: Relative change in temperature (K) from inlet to outlet of the turbine at varying air mass fraction values and inlet temperatures

Figure 14 shows the difference in the efficiency estimations based on isothermal and isentropic expansions. The isothermal efficiency is less than the isentropic efficiency for test conditions at a 0.05 air mass fraction. As a result, the apparent increase in the BEP calculated using an isentropic expansion model may not represent an actual improvement in performance. However, the change in slope due to changes in the two-phase flow interaction in nozzle itself at different pressures and low air mass fraction values can be further studied to properly characterize the performance shift.


Figure 14: Isothermal efficiency vs isentropic efficiency at a 0.05 air mass fraction and 0.483 MPa inlet pressure

The data indicate a systematic change in the performance as a function of two-phase flow. There exists the potential for characterizing and quantifying the performance by a relation that is universal to all two-phase flow conditions similar to the turbine affinity laws for single-phase flow. Further analysis is in progress to develop the scaling laws based on the current data set.


Effects of Backpressure

Investigations of the Fukushima events [26] reported that the RCIC system shutdown in Unit 3 on March 12, 2011 occurred as the result of a likely protective trip caused by high turbine exhaust pressure. To understand the effects of varied backpressure, the turbine exhaust pressure was increased by throttling the outlet control valve. Figure 15 shows the torque and efficiency at an 31

inlet pressure of 0.48 MPa, 0.70 air mass fraction and different backpressures. Systematic degradation in the performance was observed with increases in the backpressure. The efficiency envelope is narrowed with the BEP moving towards lower rotational speeds, indicating increased momentum losses. Further investigation is in progress to characterize this degradation as a function of inlet pressure, air mass fraction and backpressure.

Figure 15: Effects of backpressure on turbine performance for 0.48 MPa Inlet Pressure and 0.70 air mass fraction


Turbo-Pump Testing

Degradation in the turbine input due to change in air mass fraction will directly affect the pump power input in a turbine-driven pump. The test rig was modified, and the dynamometer was replaced by a multistage centrifugal pump (Dayton 5UXF5). As shown in Figure 1616, the pump was coupled to the turbine shaft. Rosemount pressure sensors were installed to measure the pump inlet and outlet pressures. The pump load was controlled by a control valve. A Coriolis flow meter was installed on the outlet line to measure the pump flow rate. The turbine was subjected to air mass fraction values from 1.0 to 0.3 at an inlet pressure of 0.45 MPa and atmospheric backpressure. 32

Since the pump flow rate and head output will affect the total power input, the effect of changing the pump flow rate on rotational speed is evaluated.


Figure 16: Schematic of turbo-pump test rig, highlighted in the dotted box 34

Figure 17: Variations in rotational speed with changes in the flow rate of the pump attached to the turbine shaft

For standard centrifugal pumps, the required shaft power (or torque) initially increases with increases in pump flow rate for given rotational speed. This typically continues beyond the best efficiency point (BEP). Further increases in flow rate beyond BEP results in a trend of decreasing shaft power. This is consistent with the pump’s characteristic curve. As shown in Figure 17, for all the MFs at the turbine inlet, the rotational speed initially decreases to compensate for the increase in pump torque. Further increases in the pump flow rate causes a recovery of the rotational speed due to decreased pump torque. To simplify the representation of pump performance, the rotational speeds for each mass fraction are averaged. Figure 18 shows the pump performance for these turbine inlet conditions. The turbine output (rotational speed) and consequently pump head degrades with decreases in the air mass fraction at the turbine inlet. This degradation seems systematic with the rate of pump head decrease increasing with decreases in the air mass fraction at the turbine inlet. This demonstrates how changes in turbine inlet conditions can affect the pump 35

performance, which in turn will affect the cooling water injection flowrate in a nuclear power plant. Further investigation is in progress to develop component-level prediction models to characterize the two-phase flow performance of Terry turbines.

Figure 18: Performance of the pump as a function of air mass fraction at the turbine inlet (inlet pressure of 0.45 MPa and atmospheric exhaust pressure)

Conclusions Single-phase gas-only inlet conditions: The torque decreases linearly as the rotational speed increases for constant inlet pressures; the torque, operational range, and BEP increase as the inlet pressure increases. The BEP moves towards higher speeds with increases in inlet pressure, indicating an expansion of the operational range is possible.


Two-phase inlet conditions: The torque and efficiency decrease with the BEP moving towards lower speeds as the air mass fraction decreases for constant inlet pressures. As the inlet pressure increases, the degradation in the BEP is reduced. Additionally, greater pressures reduced the momentum losses associated with two-phase interactions and increased the sonic speed while also reducing the deviation in the BEP with rotational speed. The turbine performance decreases systematically with increases in backpressure. Pump coupled to the Turbine: The degradation in the turbine performance as a function of decreasing air content reduces pump output performance. Systematic degradation in the turbine performance as the consequence of decreasing its inlet gas mass fraction can explain the turbine behavior during self-regulating operations of the RCIC system. This lays the foundation for further characterization work for Terry turbines operating under two-phase conditions.

Abbreviations BDBE

Beyond design-basis event


Best efficiency point


Boiling water reactor


Extended loss of AC power


Full scale output


Graphical user interface


Gas volume fraction


Gas mass fraction


Main steam isolation valve



Piping and instrumentation diagram

PID control

Proportional Integral Derivative control


Probabilistic risk assessment


Pressurized reactor water


Reactor pressure vessel


Reactor core isolation cooling




Station blackout

Nomenclature c

Sonic speed




Pressure difference across the turbine




Isentropic index, 1.4


Input power


Inlet pressure


Liquid volumetric flow rate






Volumetric flow rate of gas 𝑊𝑠ℎ𝑎𝑓𝑡

Shaft output work


Subscripts g, gas,

Gas Properties


Inlet Condition Properties


Outlet Condition Properties

l, liquid

Liquid properties

Acknowledgements The authors would like to thank The Institute of Applied Energy (Japan), the Japanese Ministry of Economy, Trade, and Industry (METI), the US Department of Energy, Idaho National Laboratory, Sandia National Laboratories, and the US Boiling Water Reactor Owners Group (BWROG) for financial and technical support. The authors also would like to thank Gavin Lukasik and Carl Johnson from the Turbomachinery Lab for their assistance with rig setup and experimental testing. The authors would like to thank Dr. Nobuyoshi Tsuzuki and Dr. Douglas Osborn for their technical support. Revak-Keene Turbomachinery Services of La Porte, TX is most gratefully acknowledged for turbine refurbishment services and numerous invaluable communications about Terry turbine operations, maintenance, and past experience.


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Highlights 1. A Terry ZS-1's performance was tested for gas mass fractions from 1.0 to 0.05. 2. Terry turbines show a strong adaptability under two-phase flow operating conditions. 3. The expansion of two-phase mixtures changes from isentropic to near isothermal as the air mass fraction varies from 1 to 0.05. 4. The turbine performance systematically degrades with increasing liquid content.