Combustion efficiency measurements and burner characterization in a hydrogen-oxyfuel combustor

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

Available online at www.sciencedirect.com

ScienceDirect journal homepage: www.elsevier.com/locate/he

Combustion efficiency measurements and burner characterization in a hydrogen-oxyfuel combustor Tom Tanneberger a, Sebastian Schimek a, Christian Oliver Paschereit a, Panagiotis Stathopoulos b,* € t Berlin, Mu¨ller-Breslau-Str.8, 10623 Berlin, Germany Chair of Fluid Dynamics, Technische Universita €t Berlin, Mu¨ller-Breslau-Str.8, Chair of Unsteady Thermodynamics in Gas Turbine Processes, Technische Universita 10623 Berlin, Germany a

b

article info

abstract

Article history:

Energy generation from renewable sources in the power sector keeps constantly

Received 12 December 2018

increasing. This raises the demand for fast and flexible large-scale storage technologies.

Received in revised form

Steam generation via stoichiometric combustion of hydrogen and oxygen within a steam

26 April 2019

cycle is a promising way to recombine both gases, which can be generated by electrolysis

Accepted 5 May 2019

utilizing excess renewable energy. At the same time, this technology could provide

Available online 8 June 2019

balancing and spinning network reserves. A crucial parameter of this approach is the combustion efficiency, since residual hydrogen or oxygen can damage downstream com-

Keywords:

ponents of the power plant steam cycle. The current paper investigates the combustion of

Oxy-fuel combustion

hydrogen and oxygen under steam diluted conditions. Flow field, mixing, flame types and

Hydrogen

combustion efficiency are assessed. The combustion efficiency measurement is very

Steam dilution

challenging in this case, as the combustor products consist mostly of pure steam and

Combustion efficiency

cannot be dried for conventional gas analysis. This is solved by an in-situ measurement

Large scale energy storage

method to quantify the combustion efficiency. Initial results of this approach are also presented in the current work. © 2019 Hydrogen Energy Publications LLC. Published by Elsevier Ltd. All rights reserved.

Introduction In order to reach the challenging commitments made at the UN Climate Change Conference 2015 in Paris [1], greenhouse gas emissions have to be reduced significantly. For that, the utilization of renewable energy sources needs to be expanded worldwide. In turn, this raises the problem of power generation fluctuations due to the variable generation of solar and wind energy. These fluctuations need to be balanced and already activate a rising share of primary and secondary control reserves, which are originally intended to

maintain network frequency and voltage in the case of outages. In Germany, the demands for positive and negative secondary reserve are expected to increase by 40% and 10%, respectively. The demand for tertiary reserve will be even higher and similar trends are expected for the European market in the near future [2,3]. Hence, energy storage and balancing technologies will very likely play a central role in the energy production and management sector of the near future. Different energy storage solutions are already running or are currently under development, but they all have certain

* Corresponding author. E-mail address: [email protected] (P. Stathopoulos). https://doi.org/10.1016/j.ijhydene.2019.05.055 0360-3199/© 2019 Hydrogen Energy Publications LLC. Published by Elsevier Ltd. All rights reserved.

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

disadvantages. Pumped hydro storage and compressed air energy storage, for example, lack of geographical availability. Batteries, on the other hand, suffer from limited load-unload cycles, small power density, and high storage costs [4,5]. Using mostly existing infrastructure, power-to-gas might be an easy solution for energy storage. Hereby, excess energy from renewable energy sources, e.g. wind and solar, is used to generate pressurized H2 from water. The H2 can be stored and converted back to energy when needed. This conversion is done today either by methanization, which is costly and complex, or by fuel cells, which are also costly and have limited capacity. A third opportunity is the direct use of H2 in combustion processes. Several projects on the combustion of hydrogen-rich fuels were initiated in the last years, for example the Advanced Hydrogen Gas Turbine Development Program by the American Department of Energy [6]. In April 2018, Obayashi Corporation and Kawasaki Heavy Industries, Ltd. reported the first successful operation of a 100% hydrogen-fueled gas turbine, which is incorporated in a combined heat and power plant with 1 MW electricity output in Kobe, Japan. The combustor is based on a dry low emission micro-mix combustion approach [7]. However, these technologies still produce NOx and CO2 emissions due to the parallel utilization of air as oxidizer and natural gas in gas mixtures. The project, part of which is the current work, aims to develop and investigate a zero-emission H2 combustion technology, that burns H2 and O2 under stoichiometric conditions. Such a combustor can be applied in two ways. In a short-term, it can be installed in steam power plants as an ondemand reheat stage. The power output of the plant is thus temporally boosted to provide network services like primary and secondary control reserves. This approach has been thermodynamically and economically assessed by Stathopoulos et al. [8] and it shows that the proposed system is a viable alternative for energy storage and primary control reserve. In the long term, fossil-fueled base load power plants will be replaced by renewables. In such a case, a H2/O2 combustor could be a part of a zero-emission power plant. While oxyfuel combustion of natural gas, coal and other fuels for post-combustion carbon capture and storage (CCS) is intensively investigated [9e14], there is only little research on cycles with internal stoichiometric combustion of H2 and O2, like the study of Cicconardi et al. [15]. In order to be utilized in a power plant, the exhaust steam of such a combustor must fulfill very challenging specifications, since it must produce pure, high-temperature steam. Considering a typical Rankine process, steam is generated in an externally fired boiler, expanded in a turbine and subsequently condensed. The resulting condensate is then pressurized and evaporated again. If the combustor in question is integrated into such a cycle [8], O2 and H2 residuals in its exhaust stream can oxidize or decarbonize downstream metal components. Moreover, these gases are non-condensable and can increase the condenser pressure (thus reducing the cycle efficiency). In the worst case scenario, these gases accumulate in the condenser and are not vented, which may lead to an explosive mixture. Hence, the combustion efficiency of such a burner is of crucial importance in order to reduce their concentration.

29753

Only a few attempts have been made to investigate the internal combustion of H2 and O2. For instance, in the scope of the Japanese WE-NET program [16,17], three H2/O2 combustors were developed and tested with the aim to integrate them in a hydrogen-fueled turbine based on the Rankine cycle [18]. The performed tests have demonstrated reliable ignition and stable operation of such combustors. However, these attempts did not manage to reach the aforementioned necessary combustion efficiency. Flame quenching near the combustor walls has been reported as the major reason for this result. A second project on the topic was conducted by a consortium of industrial partners and research institutes in Germany. This project focused on developing a H2/O2 steam generator as a steam cycle booster [19e23]. The combustor was adapted from rocket combustion technology and showed a very fast load availability, which is necessary to supply primary control reserve. Since the target combustion efficiency could not be reached, the project was not followed up by the time. However, Haidn et al. [24] concluded in 2009 that the concept is viable under today's circumstances, which coincides with the aforementioned analysis of Stathopoulos et al. [8]. Studies on a very similar combustion system have also been carried out in Russia [25]. Borozenko et al. [26] reported the measurement of very low hydrogen residuals in the produced steam but without providing thorough data on the uncertainty of their measurements. In contrast to the aforementioned projects, our approach is to apply the concept of steam-diluted (or wet) combustion. The idea is to decrease NOx and CO emissions even at high equivalence ratios, using high reactivity fuels like hydrogen [27e29]. In the case of hydrogen oxyfuel combustion, where none of these emissions are present, steam dilution is used to control the combustion temperature. Thus, the thermal stress of the combustion chamber can be reduced significantly compared to rocket combustors. This results in less structural complexity and less fatigue, which is crucial for the utilization in power plants. The model burner design in the current work builds upon an experimentally well-characterized design of a swirl-stabilized burner for steam-diluted H2/air combustion [30e33], which was equipped with additional O2 injectors. Preliminary results on a first and a second generation of that burner were reported by Schimek et al. [34] and Tanneberger et al. [35]. In the present study, the burner is further characterized in the sense of its isothermal flow field and mixing, as well as flame shape and type. Moreover, the diagnostics are extended by measurements of the combustion efficiency. Compared to conventional combustion systems, e.g. natural gas burners, the evaluation of the combustion efficiency is very challenging in the present case. The exhaust gas is pure steam with little concentrations of H2 and O2, which cannot be reliably sampled and dried for gas analysis. Instead, the residual gas concentration needs to be measured directly in the process. Even if combustion is complete (hc ¼ 100%), gas residuals can appear due to small deviations from stoichiometry. In order to tackle these challenges, the paper presents, for the first time in hydrogen oxyfuel combustion, an in-situ measurement technique for residual gas concentrations and combustion efficiency. Additionally, the current work introduces the novel concept of steam-diluted hydrogen oxyfuel

29754

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

combustion. Thereby, two individually swirled steam stages are applied in order to separate the reactants from each other and protect them from the cold combustor walls.

Experimental setup and measurement techniques The current project investigates a model combustor, which is described in the first subsection. The development of the combustor consists of several steps that include measurements in different types of test rigs. First, the flow field and mixing properties of the burner are studied in a vertical water tunnel. Subsequently, the burner stability maps, the various flame shapes and their connection to the aforementioned flow and mixing characteristics are investigated in an atmospheric combustion test rig. Finally, the combustion test rig is extended by the necessary diagnostics for the assessment of combustion efficiency.

Burner design The presented combustor design builds upon the experience from swirl-stabilized gas turbine combustors. However, here the main flow component is steam, which acts as a diluent for the chemical reaction and also defines the flow field in the combustion chamber. In order to achieve aerodynamic flame stabilization, an azimuthal component is imposed on the steam flow upstream of the combustion chamber. The current version of the burner utilizes two independently controllable swirl generators (inner and outer swirler in Fig. 1). This design has been chosen to generate an outer swirled steam jet that protects the reactants from the cold combustor walls in order to minimize its influence on combustion efficiency. The swirl generators consist of tangential slots between the plenum and two coaxial annular ducts that channel the steam into the combustion chamber. The swirl intensity of each swirl generator can be adjusted separately by varying the blockage of the tangential slots. As a result, two control variables are available to modify and optimize the flow field. The swirler blocking values for the configurations investigated in this work are given in Table 1. Please note that with increasing

Fig. 1 e Drawing of the investigated generic burner.

Table 1 e Burner configurations. Configuration I II III IV V

Inner swirler blocking %

Outer swirler blocking %

15 40 60 80 60

25 25 25 25 50

configuration number the swirl intensity is increased due to further blocking. H2 is injected as a non-swirling jet on the central axis, surrounded by the inner steam stream (fed through the inner swirl generator). The O2 is not contained in the steam but injected through 30 co-axial holes between the inner and outer steam ducts. By this arrangement, H2 and O2 are not able to form a combustible mixture near the walls of the combustor. Instead, they are mixed inside the combustion chamber and the risk of flashback is minimized.

Water tunnel test rig The isothermal PIV measurements of the flow field are carried out in a vertical water tunnel (Fig. 2) using a Nd:YAG laser with a wavelength of 532 nm and hollow silver-coated glass spheres as seeding. The water tunnel is also used to investigate the mixing properties by planar laser-induced fluorescence (PLIF). Separate injection of Rhodamine 6G dye is used to visualize first the O2 and H2 flow, respectively. The images are high-pass filtered at 590 nm to eliminate any impact of the laser wavelength on the fluorescence signal. The parameters of the water tunnel measurements are given in Table 2. The H2/O2 jet momentum ratio is approximately 0.2 for all cases. The total reactants-to-steam flow momentum ratio depends on the level of steam dilution, which further affects the flow rate and the Reynolds number. Thereby the absolute momentum of all three flows was conserved with respect to a reference combustion case with a thermal power of Pth ¼ 50 kW for comparability [36]. As a result of this condition, the Reynolds numbers of the fuel flows are not equal, but at the same order of magnitude. In terms of the mixing investigations, the water flow underestimates the molecular

Fig. 2 e Water tunnel test rig for PIV and PLIF.

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

29755

Table 2 e Boundary conditions of the water tunnel and combustion test rig with Pth ¼ 50 kW. Water tunnel Flow rates Reynolds number Momentum ratio Combustion test rig Flow rates Reynolds number Momentum ratio

H2

O2

Steam

0.3 7100 0.2

0.6 2800

1.86e7.44 10 k-40 k e

m3/h e e

1.5 6100 0.2

11.9 4100

21e105 10 k-51 k e

kg/h e e

diffusivity of the reactants, especially of the hydrogen. Thus, PLIF in the water tunnel predicts a slower and less homogeneous gas/steam mixing. However, this effect is rather small due to the strong turbulent diffusion in the shear layers.

Combustion test rig The combustion tests are conducted at atmospheric pressure in the test rig depicted in Fig. 3. A steam generator delivers super-heated steam to a plenum upstream of the burner at temperatures between 375 K and 600 K. Downstream of the burner an optically accessible combustion chamber made of fused quartz is installed, followed by a water-cooled exhaust tube. The combustion chamber has a diameter of 80 mm and a length of 200 mm. By means of an orifice at the end of the first exhaust tube section the exhaust gas is slightly pressurized to direct it to a Lambda Sensor (LS) via a probe. The probe has three inlets at different radial locations in order to generate a representative sample (see the close up in Fig. 3). In the second part of the exhaust section, the complete exhaust gas flow is directed through the measurement tube of a Tunable Diode Laser Absorption Spectroscopy (TDLAS) O2 sensor. The measurement procedure for the reacting test is shown in the flow chart in Fig. 4. The steam mass flow is set by a PIDcontrolled valve to a user-defined value. Manual needle valves are used to control the H2 and main O2 flows according to the required level of thermal power. As a result, the oxygen mass flow is set slightly below its stoichiometric value. An additional PID-controlled valve is used in the parallel, secondary oxygen line. It is set by the LS output and controls the total oxygen mass flow to reach stoichiometric conditions. This condition is described by the equivalence ratio f ¼ 1. Thereby, f is the H2/O2 ratio divided by its stoichiometric

Fig. 3 e Combustion test rig.

Fig. 4 e Schematic of the combustion efficiency measurements.

value. The combustion efficiency is assessed using the LS and the TDLAS sensor in the exhaust gas. For start-up, the TDLAS measurement section can be bypassed. Both gas sensors are described in detail in the following sections. Pictures of the flame shape were taken by an intensified camera at a frame rate of 125 Hz. A total number of 3640 frames per data point were taken. In order to cut off all light except the desired OH* chemiluminescence, the UV transparent camera optic is equipped with a band-pass filter (320/ 40 nm, 60% transmission). The images are averaged and postprocessed using an inverse Abel transformation. The boundary conditions of the reacting tests for a reference case with Pth ¼ 50 kW are given in Table 2. The fuel flow rates and Reynolds numbers change according to the actual power level. The lower mass flow of the steam is conserved for all tests, while the upper one is given by the cold blow-off limit (see Fig. 9). In the measurements of the combustion efficiency, the steam is super-heated to 573 K, which slightly reduces the Reynolds numbers. Moreover, the thermal power and steam mass flow rates are reduced for these tests. The mass flows are measured by Coriolis and swirl flow meters with an uncertainty of 0.5% of the measurement value.

Lambda sensor The Lambda Sensor (LS) is a stabilized ZrO2 oxygen sensor, which is commonly used to monitor the air-to-fuel ratio in internal combustion engines [37]. The ZrO2 ceramic is an oxygen-ion-conducting electrolyte that generates a voltage when in contact with gases of different O2 concentrations at high temperatures. It is thus heated up to 1000 K, ensuring a constant measurement gas temperature, and exposed to the measurement gas on one side and atmospheric air on the other side. Assuming complete combustion under perfectly stoichiometric conditions, there is no O2 in the measurement gas while the air has 20.9%vol O2. In this case, the sensor generates a threshold voltage of U ¼ 0:18 V. In the case of residual O2 in the exhaust, the O2 difference between both sides of the sensor and the corresponding voltage decreases. Additionally, the LS is also cross-sensitive to fuel gases. The fuel gas concentration on the measurement side of the sensor leads to a change of the inner material's charge density, which enhances the oxygen-ion-conduction and thus results in an increase of the sensor voltage above the threshold of U ¼ 0:18 V. This behavior is used to measure the concentration of H2 residuals in the exhaust gas. However, due to hightemperature pre-oxidation of the stoichiometric parts in the

29756

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

Fig. 5 e Calibration of the LS for O2 (left) and H2 (right). O2 and Eq. (2) for H2. The calculation switches automatically at the threshold of U ¼ 0:18 V. F denotes the Faraday and R the gas constant; U is the voltage and T the temperature of the LS; c1 , c2 and c3 are semi-empiric model constants, which are acquired by fitting the calibration data. 4FU RT

wO2 ¼ c1 e

U

wH2 ¼ ec2 Tþc3

Fig. 6 e Validation of the calibration of the LS and TDLAS system for 30 kg/h (only TDLAS), 40 kg/h and 50 kg/h steam.

(1) (2)

The calibration is conducted at a flow rate of 50 kg/h of super-heated steam at approximately 600 K. The test gas is injected far upstream to ensure optimal mixing prior to the sensor. The mass flow of the test gas is set using a fine metering valve with a maximum flow coefficient of cv ¼ 0:004. In order to set different O2 concentrations in the calibration procedure, the pressure loss over the valve is varied and measured by a differential pressure sensor with an accuracy of 0.1% of the measurement range (0  6 bar). The LS models are shown in Fig. 5 (left) for O2 and Fig. 5 (right) for H2. In both cases, the models are in good agreement with the calibration data.

Tunable diode laser absorption spectroscopy oxygen sensor

Fig. 7 e Distance between burner exit and recirculation zone.

measurement gas, solely the excess component, either H2 or O2, gets in contact with the sensor. Therefore, the LS voltage only depends on the concentration relative to the stoichiometric ratio. In order to quantify the deviations from stoichiometry, the relation between sensor voltage and concentration has been modeled and calibrated. The voltage of the LS is sampled with a sampling rate of fs ¼ 1 Hz and time-averaged over t ¼ 120 s. It is then converted using the semi-empirical model in Eq. (1) for

As stated before, the LS measures only residual gas concentrations relative to the stoichiometric ratio. However, O2 and H2 might both be present in the exhaust steam due to combustion inefficiencies. In order to quantify this stoichiometric component of residual gases, we measure the absolute O2 concentration. A tunable diode laser absorption spectroscopy (TDLAS) O2 sensor is used to directly measure the O2 concentration in-situ in the exhaust stream. TDLAS systems make use of the concentration-dependent line shape variations through light absorption at individual wavelengths for each species. They are well suited for measurements in harsh environments and were already deployed in fire suppression tests with large fractions of water (vapor and drops) [38,39] and in the high-temperature exhaust of the oxyfuel combustion of methane in an O2/CO2 atmosphere [14]. In our particular case, we expect very small concentrations of O2 in the exhaust gas. The commonly used direct TDLAS method struggles at low gas concentrations due to low signal-to-noise ratios [40]. Thus, the wavelength modulation spectroscopy (WMS) technique is used, which modulates the laser wavelength sinusoidally around the wavelength of the absorption line [41]. The concentration is thereby derived from the amplitude of the second harmonic of the modulation

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

29757

Fig. 8 e Contours of the combustion chamber with the flow coming from the burner exit on the left: Normalized axial flow field from the water tunnel (left) and normalized OH* chemiluminescence intensity (right) of Configuration III at P ¼ 50 kW with a dilution ratio of U ¼ 6:3 (top), U ¼ 4:3 (middle) and U ¼ 2:5 (bottom).

frequency. The applicability of this technique for O2 concentration measurements has been validated by Neethu et al. [42]. However, the accuracy and lower detection limit of such a sensor depend on the optical path length as well as the exhaust temperature and pressure. Since the evaluation of the combustion efficiency requires a high accuracy and a low detection limit, the laser beam is not directed perpendicular through the exhaust tube, which would lead to an optical pathway of only l ¼ 0:1 m. Instead, the exhaust is guided into a measurement tube parallel to the laser beam. This way, an optical path length of l ¼ 1:24 m is reached. The measurement tube is thermally insulated to provide a constant temperature in the measurement section. Under standard conditions (N2, 293.25 K, l ¼ 0:512 m), the system shows a deviation from linearity of less than 0.1 %vol O2, but at the actual conditions in high-temperature steam, the TDLAS system requires a new calibration. For this calibration, the same setup as for the LS calibration was used. Since the output signal is linear with respect to the volumetric O2 concentration, it only requires a two-point calibration (the same method is applied by the manufacturer of the sensor).

This is done for 0 and 3.2 %vol O2. The steam temperature during the calibration is somewhat lower than the exhaust temperature expected during the combustion experiments. Therefore, the temperature in the TDLAS measurement section is measured and used to correct the concentration measurement through a method integrated into the system and validated by its manufacturer. The signal of the TDLAS sensor is also sampled at fs ¼ 1 Hz and time-averaged over t ¼ 120 s. Additionally, the transmission of the laser is measured to assure correct measurements. Even at the harsh conditions of the high-temperature exhaust stream, the transmission of the laser is always larger than 70%.

Validation of the calibrations The LS and the TDLAS sensor are validated simultaneously in the combustion test rig (without a flame but at 600 K) for steam mass flows of 40 kg/h and 50 kg/h. Moreover, the TDLAS system is tested for 30 kg/h steam, where the pressure is too low to sample gas for the LS. The validation data is shown in Fig. 6 and aligns well with the O2 concentration, which is set

29758

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

Fig. 9 e Operational range (top) and exhaust temperatures (bottom) of Configuration III (left) and V (right) with swirlstabilized (dot) and jet (triangle) flames.

via the fine metering valve. The standard deviation from the set value is approximately 0.05 %mass O2 for both sensors.

Burner characterization The results are divided into two parts: The burner characterization and the evaluation of the combustion efficiency. The burner is characterized in isothermal and reacting conditions to gain insight into the flow field, the mixing quality and their connection to the observed flame shapes and positions as well as the burner operational maps.

Flow field The burner Configurations I-IV are investigated in the water tunnel for various steam flow rates in order to optimize the flow field towards a large operational range. Thereby, higher steam mass flows correspond to lower adiabatic flame temperatures (Tad ) in the combustion test rig. Since Tad also depends on the thermal power and preheating, the steam dilution ratio is chosen as the describing parameter instead. It is defined as the ratio of the steam mass flow rate to the sum of H2 and O2 mass flow rates U¼

m_ H2 O m_ H2 þ m_ O2

(3)

For sufficient steam dilution, the swirling flow in the combustion chamber causes vortex breakdown, which means that the flow direction is inverted on the central axis due to the positive pressure gradient over the area expansion. Thus, a recirculation zone is formed in the combustion chamber. The distance between the recirculation zone and the burner exit is plotted for the different burner configurations in Fig. 7. For all configurations, no recirculation zone is present for U < 5 because the tangential momentum is too weak to induce vortex breakdown due to the low steam mass flow through the swirl generator. Above this threshold, the axial position of the recirculation zone's stagnation point xVB diminishes with increasing steam dilution ratio. This means that the recirculation zone is shifted upstream, i.e. closer to the burner exit, due to the higher swirl intensity, which is caused by the increased steam mass flow rate. The difference between the several burner configurations is that the swirl intensity of the inner steam passage is increased from Configuration I to IV (Table 1). Hence, the recirculation zone moves closer to the burner exit from I to IV. In Configuration IV xVB is the smallest and constant for U > 7. But in case of lower U, the distance increases rapidly due to the high-pressure loss in the inner swirl generator at this configuration. Instead, Configuration III demonstrates the optimal trade-off between swirl intensity and pressure loss. It enables vortex breakdown at the lowest steam dilution ratio of all configurations. For this reason,

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

Configuration III is further investigated in the following subsections.

Flame shape and position According to the PIV measurements, the flow field can exhibit two fundamentally different modes. On the one hand, the high swirling jet causes an inner recirculation zone in the combustion chamber. On the other hand, the stable low swirl jet does not lead to vortex breakdown. This results in two different flame types, depending on the degree of steam dilution, for Configuration III. Fig. 8 shows a swirl-stabilized flame (top) for high and a jet flame (bottom) for low steam dilution. In this figure, the axial velocity field, obtained from water tunnel PIV measurements, and the flame shapes, obtained from normalized OH* chemiluminescence in the reacting tests, are depicted. Both contours are superposed by two-dimensional flow vectors. At U ¼ 6:3 (top row of Fig. 8) the axial flow is deflected towards the walls of the combustion chamber due to vortex breakdown, while a recirculation zone is established on the central axis. As expected, the flame is stabilized in the shear layers of the flow. At lower steam dilution, the flame is shifted downstream, while it is getting stretched in the axial and more confined in the radial direction. It is interesting to observe that even though no backflow is present at U ¼ 4:3, a divergence of the flow is still present due to swirling effects. This enables a stable transition from the swirl-stabilized to the jet flame type. The jet flame can be seen at U ¼ 2:5 in the last row of Fig. 8. In this case, the steam and fuel streams merge to a low swirling axial jet. The flame stabilizes at the contact surface between the H2 and O2 stream. This type of flame can be characterized as a diffusion flame. It has the longest stretch and the lowest spreading. The flame shape and the flame position depend mainly on the flame type but are only marginally affected by parameter variations for the swirl-stabilized mode. Over the complete swirl-stabilized range, the flame is only shifted downstream by approximately 0:15D when reducing the dilution ratio. The same shift can be observed in the recirculation zone. Thus the flame is constantly located at a position corresponding to 25% of the recirculation zone length for Configuration III, which indicates that its position mainly depends on the underlying flow field and not on the chemistry, even though the dilution affects flame properties, i.e. flame temperature and speed. These changes are compensated by the flame stabilization in the strong shear layers of the recirculation zone.

Operational map In order to quantify the boundaries of the two flame types, an operational map is recorded for Configuration III with thermal powers of P ¼ 20  70 kW. In the top row of Fig. 9, swirlstabilized and jet flames are marked with dots and triangles, respectively. A slow but stable transition is observed between the two flame types. The swirl-stabilized flame coincides with the existence of vortex breakdown, which requires a certain swirl number. The swirl number gives the ratio of the flow's tangential to axial momentum and is a function of the steam dilution. This means that the transition between the two flame types should occur at a constant U for a constant burner

29759

configuration. However, Fig. 9 shows that it also depends on the thermal power. For higher power levels the transition is shifted towards lower U, which means that due to the heat release of the flame a lower swirl number is necessary to enable vortex breakdown in this flow regime. Thus the swirlstabilized flame regime is enlarged with increasing thermal power and larger Tad can be reached with a swirl flame. On the right-hand side, Fig. 9 additionally shows the stability map of Configuration V. This configuration features a larger swirl intensity. Due to the increased azimuthal momentum, vortex breakdown in the combustion chamber is already enabled at lower steam dilution ratios. This promotes the swirl-stabilized flame to appear at even higher flame temperatures. The exhaust temperatures, measured downstream of the fused quartz combustion chamber, are plotted in the bottom row of Fig. 9. The exhaust temperatures are decreasing linearly with steam dilution in the swirl-stabilized range (dots). The gradient of temperature is slightly steeper for Configuration V, leading to higher exhaust temperatures at the point of transition between the jet and the swirl-stabilized flame. On the other hand, the cold blow-off of Configuration V already appears at a lower steam dilution ratio.

Mixing quality In order to achieve a high combustion efficiency, fast and complete mixing of the reactants is of high importance. The mixing process is analyzed in the water tunnel at the same conditions as the PIV measurements. Since only one flow can be visualized at a time, the mixing of O2 and H2 is investigated in separate experiments, keeping all other parameters constant. The instantaneous mixing of both gases with steam is depicted in Fig. 10 (right) for low and Fig. 10 (left) for high steam dilution. There is a clear difference in the evolution of the mixing process in both cases. At low steam dilution, the H2 injection (top row) acts like a round jet, resulting in poor mixing with the surrounding steam. The O2 jets are directed inwards to the central axis where they interact with the H2 jet at approximately x ¼ 0:5D. In the downstream part of the combustion chamber the reactants, especially the O2 fraction, show high local fluctuations. In the highly diluted case, vortex breakdown appears, changing the mixing behavior completely. The H2 jet penetrates less into the combustion chamber and spreads much more, leading to faster mixing with the surrounding steam. The O2 jets are pointing outwards and are observed to break up in the shear layers of the recirculation zone. Due to the recirculation zone, the mixture cannot be convected downstream on the central axis but is guided towards the walls of the combustion chamber. Thus, the local fluctuations in the region of the recirculation zone are low compared to the jet flame case. In the combustion test rig, this region is filled with hot exhaust gas (i.e. steam), which is recirculated to the flame zone. Both mixing processes are evaluated in the combined mixture fraction CMix , which is defined as shown in Eq. (4), where CO2 is the mixture fraction of O2 in the water tunnel, i.e. the Rhodamine concentration in the combustion chamber normalized by the concentration at the injection. The same applies to the H2 mixture fraction CH2 .

29760

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

Fig. 10 e Instantaneous mixing fractions in the combustion chamber of H2 (top) and O2 (bottom) of Configuration III for low (left) and high (right) steam dilution.

CMix ðx; yÞ ¼ CO2 ðx; yÞ þ CH2 ðx; yÞ

(4)

In order to sum up the aforementioned results, CMix is shown in Fig. 11, superposed by the stagnation line, i.e. the contour line of zero axial velocity, and a contour of the Abeldeconvoluted OH* intensity. The time-averaged mixing field is in good agreement with the shape of the stagnation line of the mean flow field in the highly diluted case. The flame is compact in axial direction and located in the shear layer of the inner recirculation zone. In the jet flame case, the flame is much longer and stabilized in the mixing layer of the reactants. Please note, that there is only low OH* intensity on the central axis where the H2 concentration is expected to be high. The axial development of the mixing process is quantified by the spatial and temporal unmixedness, defined in Eq. (6) and Eq. (7).   s20 ¼ 〈CMix ðx; yÞ〉 , 1  〈CMix ðx; yÞ〉

Us ðxÞ ¼

 2 〈 CMix ðx; yÞ  〈CMix ðx; yÞ〉 〉 s20

(5)

(6)

Ut ðxÞ ¼

 2 〈 CMix ðx; yÞ  〈CMix ðx; yÞ〉 〉 s20

(7)

where ð:Þ denotes the temporal average and 〈:〉 the spatial average in y direction. Fig. 12 is showing a lower spatial unmixedness for the swirl flame over the whole combustion chamber and lower temporal mixture fluctuations in the downstream part, which indicates a better mixing of the reactants in this case.

Conclusions on the burner characterization In conclusion, two different flame types exist with the same burner configuration by changing the steam dilution ratio. The swirl-stabilized flame results in faster and more homogeneous mixing compared to the jet flame. At the same time, the swirling flame is very compact in the axial direction and located upstream of the point where ideal mixing is reached. Due to incomplete mixing, local deviations from the stoichiometric fuel/oxidizer ratio can appear, leading to pockets of unburned gas. If these are convected away from the flame zone into colder regions, they cannot be converted into steam.

Fig. 11 e Time-averaged mixture fraction CMix for Configuration III with the contour line of zero axial velocity (white) and the flame shape (black).

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

29761

Fig. 12 e Spatial (left) and temporal (right) unmixedness for Configuration III.

This would impair the combustion efficiency especially for high dilution rates, where the flame and exhaust temperatures are already low. The jet flame is more stretched in axial direction and also covers regions, where the spatial unmixedness saturates. In the radial direction, the jet flame is very confined with low OH* intensity on the central axis, which could lead to unburned reactants. Moreover, the hot exhaust is not recirculated, so unburned fuel can easily escape the combustion zone. However, for a constant burner configuration, the exhaust temperatures are much higher for the jet flames due to the lower steam dilution (Fig. 9). This is expected to support the fuel burnout and, hence, improve combustion efficiency. The lower flame temperature of the swirl-stabilized flame is compensated by the recirculation of the hot exhaust gas, also supporting the fuel burnout. The resulting combustion efficiency is measured and discussed in the following section.

Combustion efficiency In the following subsections, first, the measurement procedure is explained in detail and second, the combustion efficiency is discussed. In Fig. 9 the operational maps of two configurations are shown. The analysis of the combustion efficiency focuses only on Configuration V because it features a larger range of swirl-stabilized combustion.

Measurement procedure of the combustion efficiency As stated before, gas residuals in the exhaust can exist due to two factors, i.e. combustion inefficiencies and equivalence ratio deviations from stoichiometry. First, the deviation from stoichiometric conditions is measured by the LS as the mass fractions wO2 ;LS and wH2 ;LS . The voltage of the LS is converted to mass fraction with Eq. (1) and Eq. (2) and subsequently timeaveraged for t ¼ 120 s in each data point. Thereby also temporary appearances of O2 or H2 are recognized, which was observed close to stoichiometric conditions. Thus one timeaveraged measurement point can contain both, H2 and O2 residuals, even though the LS only measures one or the other at a given time. The obtained mass fractions are shown in Fig. 13 as a function of the equivalence ratio, which is varied manually here in order to demonstrate its influence on the gas

residuals. As expected the LS detects O2 residuals under lean and H2 residuals under rich conditions. The total amount of O2 in the exhaust gas is measured by the TDLAS sensor in terms of the volumetric concentration fO2 . The mass fraction is then calculated by Eq. (8). The difference between the O2 mass fraction from the TDLAS and the LS yields the fraction of unburned stoichiometric O2. Based on this value, the fraction of unburned H2 can be directly derived using the stoichiometric ratio of 0.125. The mass fractions of H2 in the exhaust gas is then calculated via Eq. (10) and shown in Fig. 14. wO2 ;ex ¼

fO2 ;TDLAS ,rO2 fO2 ;TDLAS ,rO2 þ fH2 O ,rH2 O


1 1 wH2 ;ub ¼ wO2 ;ub ¼ wO2 ;ex  wO2 ;LS 8 8 wH2 ;ex ¼ wH2 ;LS þ wH2 ;ub

(8)

(9) (10)

For low steam dilution, the appearance of O2 in the exhaust gas at lean is evident as expected. The same counts for H2 in the rich regime. In contrast, at high dilution, excess O2 is also present at rich conditions, indicating an insufficient combustion efficiency. The combustion efficiency is defined in Eq. (11) as the ratio of burned stoichiometric H2 to the total amount of H2: hc ¼

m_ H2  wH2 ;ub ,m_ ex m_ H2

(11)

This way, the combustion efficiency is corrected by any deviation from stoichiometric combustion. Thus, even for divergent equivalence ratios, a high combustion efficiency can be reached. In Fig. 15 the combustion efficiency is shown for variously set equivalence ratios. While for low dilution the efficiency varies only slightly with the equivalence ratio, there is a clear drop at rich and stoichiometric conditions for high dilution ratios. However, even though the combustion efficiency is not necessarily impaired at rich or lean conditions, a deviation from stoichiometric conditions causes residual concentrations of H2 respectively O2 in the exhaust (Fig. 14). Consequently, the combustion efficiency must be investigated at controlled stoichiometric conditions, which is done in the following subsection. Fig. 16 (left) shows the combustion efficiency of burner Configuration V as a function of the steam dilution ratio. It is

29762

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

Fig. 13 e Mass fractions of the deviation from stoichiometric conditions measured by the LS for 3 levels of steam dilution.

Fig. 14 e Mass fraction of O2 in the exhaust measured by the TDLAS (left) and calculated for H2 (right) with 3 levels of steam dilution.

observed that the combustion efficiency increases with thermal power especially for high steam dilution. Steam dilution lowers the adiabatic flame temperature and is leading to a drop in efficiency above a certain level. At a steam dilution ratio above U ¼ 8 cold flame blow-off appears. The main parameter describing the decay of combustion efficiency is the exhaust temperature, as shown in Fig. 16 (right), where the efficiency curves coincide for different thermal power levels when plotted against the exhaust temperature. The efficiency

is increasing with exhaust temperature up to 1100 K and saturates afterwards close to hc ¼ 100 %. Within this saturated range, the flame is slowly changing its shape from a jet to swirl-stabilized flame as described in section Flame shape and position. However, this transition seems to have no significant impact on the combustion efficiency. In Section Conclusions on the burner characterization, the advantages and disadvantages of both flame types are summed up regarding flow field, flame shape, flame position and mixing in relation to the combustion efficiency. Considering the efficiency measurements, it seems that negative and positive features of each flame type cancel each other for Configuration V, as long as the flame temperature is high enough. The lowest measured residual gas concentrations in the exhaust were wO2 ;ex ¼ 0:21 %mass O2 and wH2 ;ex ¼ 0:03 %mass H2 at Pth ¼ 30 kW and a steam dilution of U ¼ 5.

Combustion efficiency at stoichiometric conditions

Fig. 15 e Combustion efficiency of Configuration V with 3 levels of steam dilution.

Please note that in contrast to section Operational map where saturated steam was used, the steam for the efficiency measurements is super-heated to avoid condensation in the measurement equipment. However, the super-heated test rig is much more limited in thermal power and steam mass flow rate than the saturated steam test rig. Hence, not all points of the operational map in Fig. 9 can be evaluated regarding their combustion efficiency.

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

29763

Figure 16 e Combustion efficiency of Configuration V under stoichiometric conditions.

Conclusions In the present study, results of a fundamental investigation of a steam-diluted H2/O2 combustion system are presented. The results focus especially on the combustion efficiency, which plays an important role for the industrial application of such a combustion system. First, the model burner is characterized in general by the means of PIV, PLIF and OH* chemiluminescence. Two distinct flame types are established, depending on the underlying flow field. High steam dilution ratios are leading to vortex breakdown in the combustion chamber. This enables a swirl-stabilization of the flame, which is therefore located in the shear layers between the central and outer recirculation zones. Its shape is radially distributed but axially focused. On the other hand, at low steam dilution ratios, the tangential momentum of the steam is not sufficient to generate vortex breakdown. The flow field then consists of a H2 jet in a low swirling steam environment enriched by co-axial O2 jets. Consequently, a typical jet flame with dilution flame characteristics is established. It is radially more confined and stretched in the axial direction. The mixing investigation shows a more homogeneous distribution of the reactants in the swirl-stabilized case, but the flame is located upstream of the point of perfect mixing. Nonetheless the swirl-stabilized flame profits from the recirculation of the exhaust gases. The paper presents operational maps for two burner configurations, showing the transition from the jet to the swirl-stabilized flame. A general rule of thumb is that a higher swirl intensity shifts the transition between the two flame regimes towards lower steam dilution ratios. Moreover, the onset of transition depends on the thermal power level of the flame. This is somehow unexpected since it means that an increased heat release reduces the swirl number that is necessary to enable vortex breakdown. Thus, a higher adiabatic flame temperature can be reached in the swirl-stabilized mode at increased thermal power levels. An in-situ measurement setup for the evaluation of the combustion efficiency of a steam-diluted H2/O2 flame is developed, using a lambda sensor and a tunable diode laser absorption spectroscopy O2 sensor. The measurements show, that the exhaust temperature is the governing parameter for the combustion efficiency in this particular burner configuration. For exhaust temperatures Tex > 1100 K the combustion efficiency is constantly above hc ¼ 95 % and close to 100% for high thermal

power. At lower exhaust temperatures, the combustion efficiency drops significantly. This boundary corresponds to U  6 and should be avoided in technical processes. For the first time, the jet and swirl-stabilized flame types are compared in terms of combustion efficiency. It shows that the combustion efficiency does not change significantly during the transition between both types. However, the comparison of both flame types is limited by the fact that (for a given burner configuration) their combustion efficiency could only be measured at different steam dilution ratios. Thus, the jet flame exhibits a much higher power density and exhaust temperature than the swirl-stabilized flame. In order to allow for a fair comparison regarding their efficiency, they have to be measured at the same dilution ratios. This will be incorporated in future studies, where more configurations of the burner will be investigated in an optimized setup. The steam dilution of the flame allows for a precise control of the flame temperature. The most relevant range for technical applications is a ratio of U ¼ 4  6, leading to an adiabatic flame temperature of approximately 1500 K. At a sufficiently high thermal power (Pth  30 kW), this range is covered by a stable, swirl-stabilized flame, which reaches a combustion efficiency of nearly 100%. In conclusion, the results show that swirl-stabilized hydrogen oxyfuel combustion is suitable for technical applications even in closed cycles, since the reactants are almost completely converted.

references

[1] United nations framework convention on climate change. Paris Agreement: UN Climate Conference; 2015. [2] NTSO-E. Scenario outlook and adequacy forecast. Tech Rep 2015. [3] MacDonald M. Impact assessment on European electricity balancing market. Tech Rep 2013. [4] Kousksou T, Bruel P, Jamil A, El Rhafiki T, Zeraouli Y. Energy storage: applications and challenges. Sol Energy Mater Sol Cell 2014;120:59e80. [5] Luo X, Wang J, Dooner M, Clarke J. Overview of current development in electrical energy storage technologies and the application potential in power system operation q. Appl Energy 2015;137:511e36. [6] Bancalari E, Chan P, Diakunchak IS. Advanced hydrogen gas turbine development program. In: 23rd international Pittsburgh coal conference, Pittsburgh; 2006.

29764

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 4 4 ( 2 0 1 9 ) 2 9 7 5 2 e2 9 7 6 4

[7] Tekin N, Ashikaga M, Horikawa A, Funke H. Enhancement of fuel flexibility of industrial gas turbines by development of innovative hydrogen combustion systems. Gas for Energy 2018;(2):18e23. [8] Stathopoulos P, Sleem T, Paschereit CO. Steam generation with stoichiometric combustion of H2/O2 as a way to simultaneously provide primary control reserve and energy storage. Appl Energy 2017;205(April):692e702. [9] Allam R, Martin S, Forrest B, Fetvedt J, Lu X, Freed D, Brown GW, Sasaki T, Itoh M, Manning J. Demonstration of the Allam Cycle : an update on the development status of a high efficiency supercritical carbon dioxide power process employing full carbon capture. Energy Procedia 2017;114(November 2016):5948e66. [10] Wall T, Liu Y, Spero C, Elliott L, Khare S, Rathnam R, Zeenathal F, Moghtaderi B, Buhre B, Sheng C, Gupta R, Yamada T, Makino K, Yu J. An overview on oxyfuel coal combustiondstate of the art research and technology development. Chem Eng Res Des 2009;87(8):1003e16. € cz N, Assadi M, Bai X. Characteristics of [11] Liu C, Chen G, Sipo oxy-fuel combustion in gas turbines. Appl Energy 2012;89(1):387e94. [12] Mathekga HI, Oboirien BO, North BC. A review of oxy-fuel combustion in fluidized bed reactors. Int J Energy Res 2016;40(7):878e902. [13] Ilbas M, Bektas A, Karyeyen S. Effect of oxy-fuel combustion on flame characteristics of low calorific value coal gases in a small burner and combustor. Fuel 2018;226:350e64. [14] Bu¨rkle S, Becker LG, Dreizler A, Wagner S. Experimental investigation of the flue gas thermochemical composition of an oxy-fuel swirl burner. Fuel 2018;231(March):61e72. [15] Cicconardi S. Steam power-plants fed by high pressure electrolytic hydrogen. Int J Hydrogen Energy 2004;29(5):547e51. [16] Mitsugi C, Harumi A, Kenzo F. WE-NET: Japanese hydrogen program. Int J Hydrogen Energy 1998;23(3):159e65. [17] Hijikata T. Research and development of international clean energy network using hydrogen energy (we-net). Int J Hydrogen Energy 2002;27(2):115e29. [18] Bannister RL, Newby RA, Yang W-C. Final report on the development of a hydrogen-fueled combustion turbine cycle for power generation. In: ASME 1998 international gas turbine and aeroengine congress and exhibition. American Society of Mechanical Engineers; 1998. [19] Beer S, Haidn O, Willms H. Control of stoichiometric propellant supply in an h2/o2-spinning reserve system. In: da Grac¸a Carvalho M, editor. Combustion technologies for a clean environment : selected papers from the proceedings of the first international conference. vol. 1. Vilamoura, Portugal: Energy, combustion and the environment; 1991. p. 825e38. September 3 - 6, 1991. [20] Sternfeld HJ, Heinrich P. A demonstration plant for the hydrogen/oxygen spinning reserve. Int J Hydrogen Energy 1989;14(10):703e16. [21] H. J. Sternfeld, Capacity control of power stations by O2/H2 rocket combustor technology, Acta Astronom 37. € hlke K, Haidn O. Spinning reserve system based on H2/O2 [22] Fro combustion. Energy Convers Manag 1997;38(10e13):983e93. € hlke K, Carl J, Weingartner S. Improved [23] Haidn OJ, Fro combustion efficiency of a H2/O2 steam generator for spinning reserve application. Int J Hydrogen Energy 1998;23(6):491e7. € kalp I. Clean smart grid: primary [24] Haidn OJ, Davidenko D, Go frequency control applying H2/O2 rocket combustor technology. In: 7th international energy conversion engineering conference; 2009. [25] Malyshenko SP, Gryaznov AN, Filatov NI. High-pressure H2/ O2-steam generators and their possible applications. Int J Hydrogen Energy 2004;29(6):589e96.

[26] Borzenko VI, Schastlivtsev AI. Efficiency of steam generation in a hydrogen-oxygen steam generator of kilowatt-power class. High Temp 2018;56(6):927e32. [27] Dryer F. Water addition to practical combustion systems concepts and applications. In: Symposium (international) on combustion, vol. 16. Elsevier BV; 1977. p. 279e95. [28] Bartlett M. Developing humidified gas turbine cycles. Ph.D. thesis. KTH, Chemical Engineering and Technology; 2002. € ke S, Fu¨ri M, Bourque G, Bobusch B, Go € ckeler K, [29] Go Oliver Kru¨ger, Schimek S, Terhaar S, Paschereit CO. Influence of steam dilution on the combustion of natural gas and hydrogen in premixed and rich-quench-lean combustors. Fuel Process Technol 2013;107:14e22. [30] T. G. Reichel, S. Terhaar, O. Paschereit, Increasing flashback resistance in lean premixed swirl-stabilized hydrogen combustion by axial air injection, J Eng Gas Turbines Power 137 (7). € ckeler, O. Paschereit, Investigation of lean [31] T. G. Reichel, K. Go premixed swirl-stabilized hydrogen burner with axial air injection using OH-PLIF imaging, J Eng Gas Turbines Power 137 (11). [32] Tanneberger T, Reichel TG, Kru¨ger O, Terhaar S, Paschereit CO. Numerical investigation of the flow field and mixing in a swirl-stabilized burner with a non-swirling axial jet. In: Proceedings of ASME turbo expo 2015. ASME; 2015. [33] Stathopoulos P, Kuhn P, Wendler J, Tanneberger T, Terhaar S, Paschereit CO, Schmalhofer C, Griebel P, Aigner M. Emissions of a wet premixed flame of natural gas and a mixture with hydrogen at high pressure. J Eng Gas Turbines Power 2016;139(4). 041507. [34] Schimek S, Stathopoulos P, Tanneberger T, Paschereit CO. Blue combustion: stoichiometric hydrogen - oxygen combustion under humidified conditions. In: Proceedings of ASME turbo expo 2015. Turbine Technical Conference and Exposition; 2015. [35] Tanneberger T, Schimek S, Kossatz M, Paschereit CO, Stathopoulos P. Development of a swirl-stabilized H2/O2 combustion system under humidified conditions. In: Digital proceedings of the 8th european combustion meeting. ECM 2017; 2017. p. 1618e23. [36] Lacarelle A. Modeling, control, and optimization of fuel/air mixing in a lean premixed swirl combustor using fuel staging to reduce pressure pulsations and NOx emissions. Ph.D. € t Berlin; 2011. thesis. Technische Universita [37] Brailsford A, Yussouff M, Logothetis E. A first-principles model of the zirconia oxygen sensor. Sensor Actuator B Chem 1997;44:321e6. [38] Awtry AR, Fleming JW, Ebert V. Simultaneous diode-laserbased in situ measurement of liquid water content and oxygen mole fraction in dense water mist environments. Optic Lett 2006;31(7):900. [39] Schlosser HE, Wolfrum J, Ebert V, Williams BA, Sheinson RS, Fleming JW. In situ determination of molecular oxygen concentrations in full-scale fire-suppression tests using tunable diode laser absorption spectroscopy. Proc Combust Inst 2002;29(1):353e60. [40] Silver JA. Frequency-modulation spectroscopy for trace species detection: theory and comparison among experimental methods. Appl Opt 1992;31(6):707. [41] Kluczynski P, Gustafsson J, Lindberg  AM, Axner O. Wavelength modulation absorption spectrometry e an extensive scrutiny of the generation of signals. Spectrochim Acta B Atom Spectrosc 2001;56(8):1277e354. [42] Neethu S, Verma R, Kamble S, Radhakrishnan J, Krishnapur P, Padaki V. Validation of wavelength modulation spectroscopy techniques for oxygen concentration measurement. Sensor Actuator B Chem 2014;192:70e6.