Combustion chamber design and performance for

Fuel Processing Technology 166 (2017) 258–268

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Fuel Processing Technology journal homepage: www.elsevier.com/locate/fuproc

Research article

Combustion chamber design and performance for micro gas turbine application Ibrahim I. Enagi a, K.A. Al-attab b, Z.A. Zainal c,⁎ a b c

Department of Mechanical Engineering, School of Engineering Technology, Federal Polytechnic, P.M.B. 55, Bida, Niger State, Nigeria Department of Mechanical Engineering, Faculty of Engineering, Sana'a University, Sana'a, Yemen School of Mechanical Engineering, Universiti Sains Malaysia, Engineering Campus, 14300 Nibong Tebal, Penang, Malaysia

a r t i c l e

i n f o

Article history: Received 23 January 2017 Received in revised form 31 May 2017 Accepted 31 May 2017 Available online 19 June 2017 Keywords: Micro gas turbine Turbocharger Combustion chamber Flame holder ANSYS-FLUENT

a b s t r a c t Micro-gas turbines (MGT) are small-scale independent and reliable distributed generation systems that offer potential for saving energy and reducing carbon monoxide (CO) emissions. They are expected to play a vital role in future energy supplies for remote locations with or without grid connections. In this paper, a design and development of a combustion chamber for micro-gas turbine was performed by SOLID-WORKS and computational fluid dynamics (CFD) ANSYS-FLUENT simulation software. Different chamber geometries were used to simulate with species transport and non-premixed combustion models to determine the optimum chamber design. The best chamber geometry adopted after optimization was 50 mm flame holder diameter, 60 cm chamber height, having 4 holes of 6, 8 and10 mm with dead zone between the combustion zone and dilution zone. A two-stage MGT was developed based on vehicular turbochargers to test the chamber. The experimental test of the chamber with liquefied petroleum gas (LPG) fuel resulted in a stable combustion with CO emission below 100 ppm and turbine inlet temperature below 900 °C. © 2017 Elsevier B.V. All rights reserved.

1. Introduction Micro-gas turbine (MGT) usually produces between 25 and 500 kW of electrical power, it has minimal maintenance and operational cost, high power density and low emission. An important factor that attracts researchers to develop MGT's especially for renewable energy fuel types is that it can be operated with various kinds of fuels [1]. Micro-gas turbine in recent times is given attention for decentralised generation of renewable energy [2]. There has been a renewed interest on the MGT development and deployment on small scale distributed cogeneration (DG) and poly-generation concepts [3]. When compared with other technologies for small scale power generation, they offer numerous advantages which include, high grade heat and lower emission levels, compact size, reduced noise and vibrations, easy operations and installations [4]. MGTs are essentially of two types, the first is composed of a high speed single shaft unit carrying centrifugal compressor and radial turbine (ranges from 50,000 RPM to 120,000 RPM) on the same shaft as an electrical synchronous machine. In this design, the compressor speed remains constant at generator rated speed resulting in a significant drop in efficiency at part load [5]. The second type uses a power ⁎ Corresponding author. E-mail address: [email protected] (Z.A. Zainal).

http://dx.doi.org/10.1016/j.fuproc.2017.05.037 0378-3820/© 2017 Elsevier B.V. All rights reserved.

turbine rotating at 3000 RPM connected to a conventional generator via gear box in a split- shaft designed whereas the compressor speed varies with output load resulting in a better part load efficiency [6,7]. The developmental work in small stationary and automotive gasturbines were initiated by automotive industries in 1950′s and served as fundamental achievement in today's MGT technology [8]. Modern MGT combines the reliability of an aircraft generator with low cost automotive turbocharger [8]. The common technique used to increase the power density of internal combustion engines is turbocharging which reduces pollutant emission and fuel consumption [9]. The overall turbocharger performance and turbine power have significant effects on the thermal energy transfer and engine volumetric efficiency [10,11]. The design of a combustion chamber based on temperature homogeneity and CO emission is a critical issue in the development of the MGT. Although, some typical design and experimental studies on different micro combustor configurations were carried out in the past, there is need for further design and development to improve on fundamental issues such as low outlet temperature and CO emissions. Flame sustainability over a range of operating mass flows and air-fuel (AF) ratios in a high power density micro combustion were investigated [12,13]. A high and uniform temperature distribution along the wall of the micro combustor flame tube was achieved using a stainless steel based combustor [14,15]. Another micro-axial stainless steel based combustor was developed to improve thermal performances of MGT [16,17].

I.I. Enagi et al. / Fuel Processing Technology 166 (2017) 258–268

The applications of simulation software have rapidly increased over the last decade in improving many complex processes including combustion chambers. Computational fluid dynamics (CFD) has been extensively used as an important design tool of combustion chambers. Numerous studies have utilized CFD simulation for the design and optimization of burners [18,19] and combustion chambers in MGT [6,20– 22], aviation and industrial gas turbines [17,23]. In the present work, a combustion chamber was designed and developed using CFD ANSYS-FLUENT simulation software. Species transport and non-premixed combustion models were used to determine the optimum flame holder chamber geometries. The designed combustion chamber was investigated through optimization process and the effects of different variables and geometries on the combustion stability were studied. Combustion stability was also verified experimentally during the initial tests of a two-staged turbocharger based MGT. 2. Combustion chamber design For chamber design, a criterion for selecting suitable combustor geometry was carefully examined followed by designed calculation of the dimensions. The combustion chamber design is faced with inherent challenges such as emission control, chamber materials temperature limitations and flame stability. The steps taken in designing the chamber are: SOLIDWORKS drawing of the chamber, and simulation using ANSYS-FLUENT software. 2.1. SOLIDWORKS drawing of the chamber The combustion chamber consists of the air jacket and flame tube with 100 mm fixed height conical exit of the chamber. Chamber inlet and outlet diameters were also fixed at 50 mm diameter to match the specifications of the first stage MGT Garrett GT25. Main manipulated geometries were: chamber height, chamber diameter, flame tube diameter and combustion zones configurations on the flame tube. Flame tube diameter was varied in the range of 50–150 mm, thus, chamber diameter was varied accordingly to leave air jacket space of 25 mm, while chamber height was varied from 300 to 1000 mm. Simple flame tube geometry with fixed rows of holes of 6 mm diameter through the tube height was studied first. However, for the final flame tube optimization, three zones were identified on the flame tube: premix zone, primary combustion zone and dilution or cooling zone. Each zone has a number of rows with a certain number of holes on teach row with space between zones. The number of rows was varied from 4 to 15 for each zone and number of holes in each row was also varied from 4 to 10. Two holes diameters arrangements in respective to the zones were studied: 6, 8, 10 mm and 8, 10, 12 mm. In order to simplify the meshing and simulation processes, flame tube was built as a void inside the chamber instead of a solid body. This was achieved by creating two separate bodies (chamber and flame tube) then subtracting the flame tube body from the chamber using Combine: Subtract feature in SOLIDWORKS. 2.2. Simulation using ANSYS-FLUENT software Combustion chamber geometry was optimized using ANSYS 16.1 CFD simulations program. In the Work-bench, chamber geometry was imported from SOLIDWORKS to the Design-Modeler tool. Meshing (ANSYS ICEM CFD) tool was used to create the mesh and test its quality. After that, 3D simulation was performed on the mesh using Fluent CFD tool. Species transport combustion model was tested first followed by the non-premixed combustion model. The chamber is aimed at testing liquid and gaseous fuels with MGT, thus, LPG (butane/propane) and diesel (C12H23) fuels were tested. For this initial design simulation, no swirlers were used. However, different type of commercial diesel injectors/swirlers will be simulated and compared to the experimental test with liquid fuels in future work. Air inlet boundary conditions were

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determined based on Garrett GT25 turbocharger specifications and compressor maps. Maximum air flow rate of 0.15 kg/s at 1.4 barg was considered to achieve compressor efficiency of about 70%. Optimum chamber should achieve a complete combustion with low CO emissions and stable flame propagation inside the flame tube in a compact form. Table 1 shows the main boundary conditions for air and fuel inlets, chamber outlet and walls. 2.2.1. Species transport model For flow analysis, k-epsilon viscosity model was used since it is recommended by the FLUENT theory guide for laminar-turbulent transition and turbulent flow ranges. Reynolds stress model will be used in the future to simulate different type of commercial diesel swirling injectors to be compared with the experimental test with diesel fuel. As for the combustion, energy equations were activated (i.e. non-adiabatic combustion) with species transport model used first due to its lower computing demand compared to the non-premixed combustion model. This step will provide the initial estimation of the minimum size in which the combustion will start taking place. Diesel-air and propane-air mixture were tested separately at 20% excess air with 0.15 kg/s air flow rate. And with such high flow, turbulent combustion is expected in small geometries while larger volumes will experience laminar flows in some parts of the chamber. Therefore, both laminar Finite-Rate and the turbulent Eddy-Dissipation equations were used in the volumetric combustion model. Main geometry parameters studied with this model were the chamber height and flame tube diameter. Other variables such as chamber diameter and flame tube holes configuration were neutralized. For chamber diameter, air jacket width of 25 mm was fixed, thus, making this variable dependent on the flame tube diameter variable. As for the flame tube configuration, 8 holes of 6 mm diameter were used in each row with 10 mm distance between the rows. Thus, number of rows will increase accordingly with the chamber height. Geometries with sustainable combustion were then further optimized using radiation and non-premixed combustion models. 2.2.2. Non-premixed model To achieve stabilized combustion in MGT combustors, an improved understanding of turbulent combustion through CFD simulation is required. Moreover, the accuracy of this simulation is highly dependent on the turbulence and combustion models. The non-premixed combustion model which employs the infinitely fast chemistry assumptions

Table 1 Parameters set out in boundary conditions. Parameters

Values

Fuel inlet Temperature Pressure Mass flow rate (LPG) (20%–70% excess air) Mass flow rate (Diesel) (20%–70% excess air)

300 K 2 barg 0.0053–0.0075 kg/s 0.0056–0.0078 kg/s

Air inlet Temperature Pressure Mass flow rate

530 K 1.4 barg 0.15 kg/s

Outlet Pressure Back flow temperature

1.4 barg 600 K

Inner walls Materials Emissivity

Steel 0.5

Outer walls Materials Wall thickness Heat fluxes

Steel 6 mm −10,800 W/m2

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was therefore applied to predict the diffusion flames. Further optimization was carried out on the accepted geometries using non-premixed combustion model. Probability density function (PDF) of diesel and LPG fuels were used separately for the combustion model. LPG fuel in Malaysia contains about 70% butane, 29.5% propane and 0.5% pentane/ other components in volume [24]. Therefore similar values were set in the PDF in order to compare the results with the experimental data. Radiation P1 model was introduced to take account of radiation heat transfer inside the chamber with stainless steel wall emissivity of 0.5. Convection heat flux loses through the outer jacket walls were found to be about10800W/m2. The XY plots of temperature and CO for chamber outlet were generated to calculate average turbine inlet temperature (TIT) and mole fraction of CO in ppm. Three main simulation outputs were investigated to determine the optimum design: • CO emissions that indicated the combustion completion. • Outlet temperature that indicates the effectiveness of the dilution or cooling zone. Hence, lower temperature is desired to prevent damaging MGT turbine blades. • Flame propagation inside the flame tube. The cases with disturbed flame where flame exits the tube to the air jacket were rejected since the tube becomes surrounded by flame at some points without air cooling causing damage to the tube material. Flame shape inside the tube was also studied to prevent the direct contact between intensive and high temperature flames and tube walls.

The main geometry manipulated variables were: chamber height, flame tube diameter (chamber diameter will change accordingly) and hole zones on the flame tube. Zone arrangement on the flame tube included holes diameters, number of holes in one row, number of rows per zone and distance between zones which will be referred to as dead zones since it does not include any holes.

3. Experiment test rig 3.1. Setup A test rig is aimed at testing gaseous and liquid fuels on MGT was developed. The test rig consists of a combustion chamber designed using CFD simulation, two-stage turbocharger based MGT and exhaust gas heat recovery unit (HRU). The optimum combustion chamber design was fabricated using 6 mm thick stainless steel (SS) pipes for the outer air jacket and flame tube, while the top plate was fabricated using 10 mm thick SS plate. The optimum geometry is 600 mm in height and 152 mm inner diameter with a 100 mm height nozzle at the chamber exit. The flame holder was based on flame tube design of 690 mm in height and 50 mm inner diameter. Combustion zones contained 4 rows of holes with 8 holes in each row with largest dead zone of 350 mm between primary and dilution zones as shown in Fig. 1(a). Holes diameters for premixing, primary and dilution zones were 6, 8 and 10 mm, respectively. A 6 mm diameter LPG injector of was fabricated for the experiment based on the simulation. As for diesel injector, four sizes of commercial injectors 1–4 GPH with 30° spray angle and swirler will be simulated and compared to the experimental test with diesel in a future work. Garrett GT 25 was used on the first stage and Holset H1C for the second stage. The compressor was removed from the H1C and replaced with aluminium friction disc connected to a Teflon brake dynamometer. Exhaust gas from the second stage turbine was recovered in HRU heat exchanger for hot air production. The HRU comprises 100 annular counter-flow tubes with 6 mm diameter cupper inner tube for gas flow and 19 mm diameter mild steel outer tube for air flow. HRU design and test performance were published in earlier work [6,25]. Other auxiliary units include, oil pump, oil cooling radiator, liquid fuels pump, MGT start-up air blower and high volt DC transformer for spark plug ignition. Fig. 1(b) shows a schematic drawing for the system.

Fig. 1. (a) Combustion chamber drawing; (b) schematic drawing of the test rig.


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3.2. Procedures and measuring equipment

4.2. Chamber optimization

For the initial test of the rig, LPG fuel was used to evaluate the combustion chamber, MGT and the Teflon brake dynamometer. The optimum combustion chamber design was evaluated in terms of combustion stability, emissions and the capability of cooling zone to drop TIT down in the range of 900–1000 °C to prevent damaging turbine blades. Variety of liquid bio-fuels such as biodiesel, diesel-cooking oil blends and pyrolysis bio-oil will eventually be tested on the rig. Five type-K thermocouples were used to capture the temperature profile for the combustor, MGT and HRU. Also, a 2 barg pressure gauge was used to measure the MGT compressor pressure additional to four differential pressure gages to measure pressure drop in the combustor, first and second stage turbines and HRU. Mechanical shaft power of the second-stage turbine or the power turbine was determined using 50 mm Teflon brake disc connected to a 5 N spring scale to measure the torque. Power turbine speed was limited to 20,000 RPM to prevent overheating the Teflon brake disc and torque readings were taken at 10,000, 15,000 and 20,000 RPM. A vortex-tube cooler was used to cool the disc and a 100,000 RPM optical tachometer was used to measure the power turbine speed. Air flow rate was measured using hot-wire anemometer while LPG flow rate was measure using a 100 LPM rotameter. For MGT start-up, a ring-type 750 W air blower was used. LPG was ignited with high-volt spark plug and at 0.2 barg compressor pressure, the first-stage MGT can run in a self-sustain mode and the air blower can be shut down. The MGT was tested in the range of 0.2–0.7 barg and the measurement of air and LPG flow rates, temperature and pressure profiles and shaft brake power were taken every 0.1 barg intervals.

The combustion chamber was further optimized with the radiation P1 model and non-premixed combustion model with PDF for diesel and LPG. Chamber heights were varied from 300 mm up to 700 mm with flame tube diameter in the range of 50–100 mm. Simulation results with LPG fuel will be illustrated in this section for better comparison with the experimental work.

4. Simulation results and discussions The species transport model revealed the preliminary result of combustion for different flame tube heights and diameters. However, it did not indicate combustion completion since the model does not provide CO emissions. Hence, non-premixed combustion model with radiation model and convective heat losses from the chamber were used to determine optimum chamber geometry.

4.1. Species transport model The simulations were first performed on different 3D geometries of chamber heights and flame holder diameters using species transport combustion model. For air flow rate of 0.15 kg/s, chambers with low height below 300 mm were not able to sustain flame at all and even with the patched ignition temperature of 2000 K, temperature dropped directly back to the boundary temperature. Table 2 shows the combustion sustainability with diesel and propane fuels for chamber heights in the range of 300–700 mm. Propane fuel has proven to be easier to ignite and sustain the flame at smaller geometries. As for larger geometries, combustion was easily sustainable, however, minimal volume and surface area is preferable to reduce heat loses and manufacturing cost.

Table 2 Results obtained for species transport model. Chamber height (mm) Flame holder diameter (mm) 50 75 100 125 150

300

400

500

600

700

P P,D P,D

P P,D P,D P,D P,D

P P,D P,D P,D P,D

P,D P,D P,D P,D P,D

P,D P,D P,D P,D P,D

Note: (P) sustainable combustion with propane; (D) sustainable combustion with diesel.

4.2.1. Effect of number of rows For the flame tube zones optimization, first arrangement of 6, 8, 10 mm hole diameters for premixed, combustion and dilution zones, respectively, was fixed while changing other variables. First variable was number of rows were 4, 7 and 15 rows in each zone were tested with chamber heights in the range of 300 mm to 700 mm with 50–100 mm flame tube diameter range. It was noticed that increasing number of rows increases the flame bleeding out to the air jacked in all cases with no noticeable difference between LPG and diesel. An example temperature contours for 700 mm chamber height with 15 rows is shown in Fig. 2(a) and 7 rows for 300 mm chamber height in Fig. 2(b). On the other hand, in the 4 rows design, flame was mostly contained inside the tube even for the low chamber height (300 mm) with minor bleeding at the rich fuel premix zone as shown in Fig. 2(c) and (d) for 700 mm and 300 mm chamber heights respectively. Although the number of rows has affected the flame pattern noticeably, it did not affect combustion quality in term of maximum temperature and CO emissions.

4.2.2. Effect of holes diameters The effect of flame tube holes diameters on flame stability was investigated by increasing the diameters from 6, 8, 10 mm to 8, 10, 12 mm hole diameters arrangement as shown in Fig. 2(e) and (f). It was concluded that increasing the cross flow surface area of the holes between air jacket and combustion zone will lead to flame bleeding to the jacked. Hence, the increment in holes diameters resulted in a similar result to the addition of more rows. Another note is that the flame bleeding occurred mostly at the rich flame zone which is expected since the flame tends to propagate towards richer oxygen medium if the path has low flow resistance. Thus 4 rows of 6 mm diameter holes were sufficient to maintain rich fuel flame inside the tube.

4.2.3. Effect of flame tube diameter Chambers with low height below 600 mm did not show a good performance in terms of flame stability and CO emissions. Therefore, three chamber heights of 600, 700 and 800 mm were further compared with 25, 75 and 100 mm flame tube diameters. Results shown here in this section are for LPG PDF simulation. However, diesel fuel has shown similar trends with the different geometries, but with slightly higher temperature and CO emissions. In order to determine the optimum geometry, the main manipulated parameters were the average CO emissions and temperature at the chamber outlet. Fig. 3(a) and (b) show average CO emissions and temperature, respectively, for the three heights with LPG fuel and 70% excess air. It can be noticed that most of the geometries generated outlet temperature slightly above 1200 °C, hence, temperature was not a considerable variable for the comparison between geometries. On the other hand, flame tube diameter showed a considerable effect on CO emissions. Although, the flame tube diameter had no effect on combustion setup in terms of air-fuel (AF) ratio, it affected the velocity and turbulence profiles inside the tube. Therefore, the smaller 50 mm diameter flame tube with higher flame turbulence and better air-fuel mixing inside the tube resulted in a slight enhancement in combustion reducing CO emissions in the range of 100–150 ppm compared to 200–300 ppm for larger flame tubes.

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Fig. 2. Temperature contours for (6,8,10) mm configuration (a) 700 mm height with 15 rows; (b) 300 mm height with 7 rows; (c) 700 height with 4 rows (d) 300 mm height with 4 rows; (e) 700 mm heights with 4 rows (8,10,12) mm configuration (f) 300 mm height with 4 rows (8,10,12) mm configuration, all with LPG fuel.


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Fig. 3. (a) Average CO mole fraction for different geometries; (b) average outlet temperature for different geometries; and Average CO mole fraction for: (c) 800 mm height with and without radiation model, (d) 60 & 70% excess air, (e) dead zones A and B, all with LPG fuel.

4.2.4. Effect of radiation model Internal heat transfer inside the flame tube is mainly carried out by the forced convection by flame propagation and also natural convection causing heat to travel vertically to the top of the chamber. However, all air and fuel species as well as the internal surfaces of the flame tube are also heated by the radiation emitted from the flame. Radiation has a positive effect on combustion through additional heat to further burn the remaining unburnt CO species. Fig. 3(c) shows the effect of P1 radiation model on the average CO emissions at the chamber outlet. Flame tube diameter of 50 mm showed a superior performance compared to the higher diameter of 75 mm. Also, extending the height of the chamber up to 800 mm with the additional heat loses as surface area increases and the additional material cost did not show any combustion enhancement. Thus, further testes were carried out on 600 and 700 mm heights with 2 in. flame tube diameter.

4.2.5. Effect of excess air supplied In general, turbine engines require much higher amount of combustion excessive air compared to other thermal engines. Only a fraction of the supplied air is used for a near-stoichiometric combustion and the rest is used to dilute and cool down the combustion product gasses to prevent damaging of the turbine blades. Also, the amount of excess air is a major parameter in the thermal efficiency of a combined heat and power systems since most of the fuel input power is converted to sensible heat at the turbine exit [26]. In previous studies [27,6] a Mitsubishi TD05 turbocharger which has similar capacity to the GT25 turbocharger used here, was tested at nearly maximum flow capacity. Stable MGT operation was achieved using LPG fuel with excess air supply in the range of 30–70%. In this simulation, air flow rate was fixed at 0.15 kg/s while reducing fuel flow rate to achieve higher excess air values. The effect of excess air on CO emissions is shown in Fig. 3(d) with 60% and 70%

264


excess air. A considerable drop in average CO mole concentration at the outlet was achieved by reducing fuel flow rate while maintaining same air flow which is expected. However, in practice, excessive amount of air without controlling and canalling air to the dilution zone will result in a near lean-limit combustion. This can affect combustion stability significantly and lead to frequent flame blowout. 4.2.6. Effect of dead zones It was established earlier in the preliminary optimization of the zones geometry that a stable combustion can be achieved with 4 rows of holes with 6, 8, 10 mm holes diameters. However, the distance between zones (i.e. dead zone) also plays a considerable role in flame stability and combustion quality, but to a lesser extent. Two dead zone

configurations A and B shown in Fig. 3(e) were compared. Early introduction of dilution air (configuration A) into the non-fully established flame resulted in a considerably higher CO emissions especially in the limited chamber height of 600 mm as shown in Fig. 3(f). The length of the primary combustion zone in MGT combustion chambers is a major concern in combustion quality [28]. Extending the primary zone (configuration B) to allow most of the species to complete the reaction before introducing the dilution air has resulted in a significant drop in CO emissions for both 600 and 700 mm chamber heights. The 600 and 700 mm chamber heights did not show a considerable difference in CO emissions and chamber outlet temperature. Thus, the 600 mm chamber height, 50 mm flame tube diameter, with 4 rows of holes of (6,8,10) mm diameters, and extended dead zone between

Fig. 4. Optimum chamber geometry (a) temperature plates; (b) CO mole fraction plates; (c) stainless steel flame tube visual inspection after test; and optimum geometry simulated at the experimental conditions: (d) flame tube temperature, (e) temperature plates, all with LPG fuel.


combustion and dilution zones was chosen as the optimum chamber geometry. Using XY plot option in Fluent, average temperature and CO mole fraction for the chamber outlet were 1218 °C and 102 ppm, respectively. Fig. 4(a) and (b) show the temperature and CO distribution in two vertical XY, ZY plates. The flame was well distributed within the flame holder with minor flame bleeding at the rich mixture mixing zone. 5. Combustor performance For the initial combustor and MGT test, LPG fuel was used since it provides more flexibility and control on turbine speed and output power. The non-premixed combustion CFD simulation was repeated on the optimum geometry but at the new boundary conditions obtained from the experiments with LPG fuel at maximum turbine power. Boundary conditions were set as 0.7 barg, 0.07 kg/s air flow rate and 37% excess air in order to verify the CFD model. Fuel species were set in the PDF at 70% butane and 30% propane to match the LPG used in the experiments. The flame tube was visually inspected after the test as shown in Fig. 4(c). It was noticed the flame started to propagate right after the premixed zone causing stainless steel surface oxidization and discoloration with bluish hot stains on the flame tube walls between premix and dilution zones. On the other hand, the tip of the tube after the dilution zone holes maintained it metallic shining appearance indicating that the metal surface temperature was below 400 °C in

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which surface oxidization did not take place. Combustion has extended through the dead zone between combustion and dilution zones, while cold air jets through dilution holes has pushed the flame front away from tube walls to the center of the tube. This indicated that the dilution zone geometry was sufficient to cool down the combustion product gas and drop the turbine inlet temperature below 900 °C. Simulation temperature profile results of the flame tube inner and outer walls and chamber temperature contours are shown in Fig. 4(d) and (e). Air jacket played a significant role in cooling the flame tube walls to prevent damaging the tube. A comparison between the hot zones in simulation with the visual inspection of the flame tube showed a good agreement and explained the hot spot below the premixed zone cause by the flame leakage. The CFD model provided valuable data that were not attainable otherwise by experiment such as the flow pattern and AF ratios at each zone inside the chamber. AF ratios varied significantly from the center of the flame tube towards the tube walls in all zones. This is because air was introduced radially towards the center of the tube where fuel is concentrated resulting in a reduction in AF ratio towards the center. At the premix zone, low AF ratio of 3 dominated most of the tube and only elevated at the air inlet holes on the tube walls. This resulted in a small combustion spots at the holes as shown by the temperature vectors in Fig. 5(a) for the first row in premix zone. Also, the flow pattern shown in the same figure reveals a slight fuel leakage towards air jacked caused by the air stream around the tube that cause eddies and lower

Fig. 5. Optimum geometry simulated at the experimental conditions with LPG fuel: (a–c) velocity vectors colored by total temperature for (a) premixed zone (b) combustion zone (c) dilution zone; (d) mole fraction of CO; (e) formation rate of thermal NOx; (formation rate of prompt NOx).

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Fig. 6. (a) MGT temperature profile; (b) chamber CO and NOx emissions.

pressure zones where air stream is separated from the tube. The size of the fuel rich zone at the center of the tube shrinks as fuel is further consumed down the tube towards the combustion and dilution zones. As for the combustion zone (rows 5–8) in Fig. 5(b), the fuel rich zone with AF ratio of 4 is easily identified by its lower temperature surrounded by flames in a rich to stoichiometric combustion (AF 10– 15). The flow pattern in the combustion zone is more stable compared to premix zone with uniform air flow through all the holes. Fig. 5(c) shows the flow pattern and temperature profile in the dilution zone. Cooling air is pushed through the larger holes of 10 mm in diameter causing a significant drop in average temperature before the turbine. In general, the flame started as rich combustion hot spots near the air inlet holes in premixed zone with low AF ratio generating high amounts of CO as shown Fig. 5(d). Rich combustion also propagated between the combustion and dilution zones as indicated by the high CO concentration in the same figure. However, CO concentration dropped significantly at the dilution zone with lean combustion and AF ratio above 15 with calculated average CO mole fraction at the outlet of 69 ppm. NOx model was implemented for the verification model with thermal, prompt NOx and reburn options. Fig. 5(e) and (f) show the generation of thermal and prompt NOx respectively. The formation of thermal NOx is mainly dependent on the temperature and the availability of free

oxygen for the reaction. Prompt NOx, on the other hand, can occur at lower temperature when free hydrocarbon radicals react with nitrogen [29]. The generation of thermal NOx was concentrated mainly at the highest temperature zone near the inner wall of the flame tube shown earlier in Fig. 4(e) where AF ratio is slightly above stoichiometric condition. As for the prompt NOx, the generation started at the rich fuel premix zone where fuel chains broke down to lighter hydrocarbon such as methane that can react with nitrogen to form NOx. The formation of prompt NOx was also detected at the combustion zone but it did not Table 3 Comparison between experimental and CFD simulation results. Parameters

Experiment

Simulation

Error (%)

Fuel composition (Vol%) Air inlet pressure (barg) Air flow rate (kg/s) Excess air (%) TIT (°C) CO (ppm) NOx (ppm)

70% butane and 29.5% propane 0.7

70% butane and 30% propane 0.7

– –

0.07 37 920–960 50–60 34–20

0.07 37 1069 69 4

– – 16.2–11.4 20–15 88.2–80


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Fig. 7. Brake power output and combustion excess air.

extend down to the dilution zone since most of the fuel was converted into CO and CO2. Average mole fraction of the total NOx formation at the outlet was below 5 ppm. The chamber and MGT were tested for extended period at 0.5 barg MGT pressure and it was noticed that TIT did not exceed 880 °C as illustrated in the MGT temperature profile in Fig. 6(a). Using low pressure domestic LPG tank, the maximum operation pressure was about 0.7 barg. Fig. 6(b) shows the experimental CO and NOx emission results during MGT start-up (0.05 barg) and up to maximum pressure of 0.7 barg at different TIT. Combustion temperature presents the main factor that governs CO production. However, since temperature was not measured experimentally at the different combustions zones, TIT was used as a direct indication of combustion intensity. CO emission was slightly above 100 ppm during MGT start-up due to the low excess air and combustion temperature with low NOx emissions below 5 ppm. However, after the start-up, compressor speed elevated rapidly providing higher air pressure and flow rate, thus, resulting in more intensive flame at the combustion zone and lower CO emission but with slightly elevated NOx emission up to 34 ppm. Table 3 shows a comparison between the results obtained from the CFD model and experiment with LPG fuel at similar operating conditions. The CFD simulation has provided good prediction of the flame distribution across the flame tube. However, temperature and CO emission values were slightly over estimated while lower values were predicted for NOx emission compared to the experimental results. 6. MGT performance A by-pass valve between the two stage turbines was used to release some of the power turbine pressure to prevent shaft over speeding when the brake is not engaged. The valve is closed during the brake power testing only to prevent Teflon brake pad from overheating with additional vortex tube cooler to cool the pad. The MGT was operated with domestic low pressure LPG tank, hence, maximum MGT pressure was 0.7 barg. Fig. 7 shows the brake power and combustion excess air at different MGT pressures. Maximum brake power around 0.55 kW at 20,000 RPM, however, actual brake power is expected to be higher due to leakage from the by-pass valve. The power turbine will be connected to a high speed alternator in the future development of the unit and by-pass valve will not be needed. 7. Conclusion The design and development of MGT combustion chamber was performed, using CAD SOLID-WORKS and ANSYS-FLUENT 16.1 CFD simulation software. Various design parameters and combustion models were

used to produce an optimum design configuration. The simulations were first performed on different 3D geometries of chamber heights and flame tube diameter using species transport combustion model. The non-premixed combustion model with laminar finite-rate and the turbulent eddy-dissipation equations was further applied with radiation P1 model to optimize the flame tube geometry. Chamber geometry of 600 mm height and 152 mm diameter and 50 mm flame tube has provided low CO emissions of 102 ppm and chamber exhaust temperature of 1218 °C. Optimum flame tube configuration included three zones: premix, combustion and dilution. Each zone consisted of 4 rows of holes of 6, 8 and 10 mm diameters respectively with 350 mm dead zone between the combustion and dilution zones. The chamber was fabricated and tested experimentally with turbocharger based two-stage MGT. LPG fuel was used for the preliminary test of the chamber. The dilution zone has proven to be effective in cooling the combustion product gasses below 900 °C without disturbing the combustion process with low CO emissions below 100 ppm. A Teflon brake dynamometer was connected to the power turbine and maximum brake power was about 0.55 kW.

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