Experimental Determination of the Efficiency and Emissions of a

Article pubs.acs.org/EF

Experimental Determination of the Efficiency and Emissions of a Residential Microcogeneration System Based on a Stirling Engine and Fueled by Diesel and Ethanol Nicolas Farra,† Tommy Tzanetakis,† and Murray J. Thomson*,† †

Department of Mechanical and Industrial Engineering, University of Toronto, 5 King’s College Road, Toronto, Ontario, Canada M5S 3G8 ABSTRACT: Renewable forms of energy, such as biofuels, have the potential to displace fossil fuels in a wide variety of applications. Further benefit can be achieved through the use of combined heat and power devices, as this is accompanied with a considerable increase in energy efficiency and lower costs associated with fuel consumption. In this study, we examined the performance of a residential microcogeneration system based on a Stirling engine and fueled by diesel and ethanol. Run on diesel, and on a lower heating value basis, the system achieved a power efficiency of 12.1%, a thermal efficiency of 73.3%, and a total efficiency of 85.4%. Powered by ethanol, the corresponding efficiencies were 11.8%, 73.9%, and 85.7%, respectively. During steady state operation, the total unburned hydrocarbon emissions for both fuels were negligible, while the particulate emissions for ethanol and diesel were found to be 0.40 mg/kWh and 0.42 mg/kWh, respectively. Emissions were extremely low, as the combustor features a continuous premixed flame that facilitates the complete burnout of already evaporated fuel. Though emissions of nitrogen dioxide, methane, formaldehyde, and acetaldehyde were also negligible for both fuels, carbon monoxide and nitric oxide emissions for diesel (71 and 67 mg/m3, respectively) were much higher than those observed for ethanol (50 and 19 mg/m3, respectively). Lower nitric oxide emission levels for ethanol were attributed to its lower flame temperature, whereas reductions in carbon monoxide emissions were likely a result of a higher degree of fuel/air mixing with ethanol, due to higher gas jet velocities of the fuel exiting the orifices of the evaporator. Lastly, parametric studies on primary engine set points, including coolant temperature and exhaust temperature, were conducted to understand their effect on engine performance.

1. INTRODUCTION The accumulation of greenhouse gases in the atmosphere is one of the primary factors known to accelerate global climate change. According to the International Energy Agency, 81.3% of the world’s energy supply in 2008 comprised of fossil fuel sources, which contributed to 99.6% of all carbon dioxide emissions.1 Thus, immediate action is required to significantly reduce emissions from their current levels in order to properly address this global issue. A proposed solution is to alleviate the effects of global climate change by complementing fossil fuels with renewable forms of energy, such as biomass-derived fuels known as biofuels. Biofuels are derived from recently dead biological materials and can potentially have zero net greenhouse gas emissions, since their feedstock consumes carbon dioxide throughout its life by photosynthesis. Although biofuels have been used throughout history, fossil fuels have ultimately replaced their use due to their relative abundance and high energy content. However, renewed interest in biofuels has emerged more recently due to an increase in the cost of fossil fuels, issues with long-term supply, fears of global climate change, and, in particular, due to the advantages in using second generation nonfood-based feedstock. 1.1. Cogeneration. One particular technology that can be used to reduce the effects of global climate change is cogeneration. Cogeneration, or combined heat and power (CHP) technology, is the production of more than one useful form of energy from a single fuel source. This mode of operation differs greatly from conventional fossil fuel-based electricity generating systems, as it involves the utilization of © 2012 American Chemical Society

both heat and power. As a result, cogeneration systems exhibit an increase in the efficiency of energy conversion, which is accompanied by a net reduction in greenhouse gas emissions and lower costs associated with fuel consumption.2 These features make cogeneration an ideal option for use in residential applications with the added benefit of on-site power consumption that avoids transmission losses. Small-scale cogeneration used in the residential sector is referred to as microcogeneration. Microcogeneration technology typically incorporates an internal combustion (IC) engine, a Stirling engine, a fuel cell system, or a microturbine system. Current commercial microcogeneration systems for single family applications produce 0.5−6.0 kW of power and 1.5−15.0 kW of heat. While the heat is utilized for space and hot water heating, the power can be used directly or it can be net metered. This means that it can be used to supplement the grid’s electrical supply. Table 1 lists existing systems along with their cogeneration technology, nominal power output, and nominal thermal output. 1.2. The Stirling Engine. Primarily used in CHP applications, Stirling engines are reciprocating external combustion engines. Unlike IC engines, combustion in a Stirling engine does not occur within the engine’s cylinders but rather in a chamber adjacent to the engine block itself. As a result, a wide range of energy sources can be used as the source Received: September 27, 2011 Revised: December 26, 2011 Published: January 27, 2012 889

dx.doi.org/10.1021/ef201468j | Energy Fuels 2012, 26, 889−900

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electrical efficiency (these issues are discussed in more detail in Section 3.1). While numerous performance studies have been conducted on Stirling engine-based cogeneration systems, only a handful have focused on small-scale cogeneration units.2,10−20 Onovwiona et al.2 have reported the emission characteristics of a cogeneration unit developed by SOLO. Featuring a natural gas burner and producing 2−9 kW of power and 8−24 kW of heat, emission levels were fairly low with 80−120 mg/m3 of nitrogen oxides (NOx), 40−60 mg/m3 of carbon monoxide (CO), and trace levels of unburned hydrocarbon (UHC) and soot emissions. An efficiency and emissions study has also been conducted by Aliabadi et al.10 on a residential Stirling enginebased cogeneration system developed by Whisper Tech. The system features a diesel fuel burner and produces 0.6−1.1 kW of power and 5.5−7.0 kW of heat. Emissions were also found to be low with 158 mg/m3 of NOx, 21 mg/m3 of CO, 0.65 mg/m3 of soot, and a trace amount of UHC emissions. The engine was also operated with biodiesel, which resulted in similar emission levels. The current study uses the same Stirling engine and is closely based on the work done by Aliabadi et al. However, there are several important contributions in this study which separate it from previous work. These include the modifications made to the fuel supply line and control system in order to accommodate ethanol as a fuel, the parametric study of exhaust gas and coolant temperature set points on engine performance, and the improvement or correction of some emissions and efficiency measurements that were originally reported by Aliabadi et al.10 1.3. Ethanol. Ethanol is a colorless and flammable alcohol that is typically produced by fermenting either starch from corn or sugar from sugar cane. Although ethanol is used in a wide variety of industrial applications, it has a number of properties that make it suitable for use as a biofuel. Advantages include its volatility and its renewable, less polluting nature compared to diesel fuel. Concerns over ethanol production are due to its relatively high cost, competition with the food supply, differing studies on its net energy gain, and the use of high-quality agricultural land. However, a distinction is made between sugar cane-based ethanol and corn-based ethanol, as sugar cane-based ethanol has substantially higher net energy yields per hectare and much lower greenhouse gas emissions associated with production.21 Alternatively, ethanol production can involve the use of biochemical processes on biomass to produce lignocellulosic ethanol. The main benefits of this fuel include its derivation from nonfood-based sources and that it can be grown on agriculturally marginal lands. Despite these benefits, lignocellulosic ethanol production must be made economically feasible through further development in various microbiological processes.22 Various studies have been conducted on the utilization of ethanol in both spark ignition and compression ignition

Table 1. Residential Microcogeneration Systems

a

manufacturer

cogeneration technology

power (kW)a

heat (kW)a

Senertec Freewatt Whisper Tech Ebara Ballard

IC engine IC engine Stirling engine fuel cell

5.5 1.2 0.8 1.0

12.5 3.3 5.5 1.5

Taken from various literature sources.2−4

of thermal energy, including conventional fossil fuel-based fuels such as gasoline or diesel, and renewable energy sources such as biomass, solar energy, and process heat. Since combustion takes place outside of the engine, this results in a well-controlled continuous combustion process. Thus, emissions from Stirling engines can be ten times lower than IC engines with catalytic converters, making them comparable to modern gas burner technology.2 Furthermore, the sealed operating chambers of the engine result in low wear and long maintenance-free operating periods. Despite their high costs, renewed development in Stirling engines is in progress because of their high level of reliability, fuel flexibility, quiet operation, and their ability to achieve high efficiency, low emissions, and good performance at partial loads.2,5 The Stirling engine operates on a closed cycle, meaning that the working fluid is enclosed within the engine’s cylinders and is thus completely independent of the combustion process. The ideal Stirling cycle consists of four reversible gas processes: isothermal expansion, constant volume regenerative heat rejection, isothermal compression, and constant volume regenerative heat addition.6 Although the ideal Stirling cycle achieves the Carnot efficiency, inefficient heat transfer, material limitations, and the presence of friction have a parasitic effect on engine performance. A variety of manufacturers have incorporated Stirling engines as a central component in their cogeneration systems. Table 2 describes recently developed systems along with their drive mechanism, working fluid, fuel type, nominal power output, power efficiency, which is based on the fuel’s lower heating value, and power-to-heat ratio. It should also be noted that most of these units are capable of operating with a range of fuels. In addition, these engines differ greatly when considering the total number of cylinders used, the mean cycle pressures of the working fluid and the specific type of drive mechanism used to couple the engine’s pistons. The considerable variation in Stirling engine design results in a wide range of systems with respect to scale and performance. For example, the Whispergen DC has a relatively low power-to-heat ratio compared to the other units. This is due to the fact that it was designed primarily as a battery charging system for boats.9 Furthermore, the unit uses relatively low pressure nitrogen as the working fluid. This is not an ideal fluid for Stirling engines and contributes to lower Table 2. Cogeneration Systems Based on Stirling Engines manufacturer

unit

drive mechanism

working fluid

fuel

power output (kW)a

power efficiency (%)a

power-to-heat ratioa

DTE Energy SOLO Sunpower Whisper Tech

ENX 55 Stirling 161 EG−1000 Whispergen DC

kinematic kinematic free piston kinematic

hydrogen helium helium nitrogen

natural gas natural gas propane diesel

55 9 1 0.8

30 24 32 12

0.60 0.25−0.36 0.55b 0.14c

a

Taken from various literature sources.2,4,7,8 bBased on 90% total efficiency and 32% electrical efficiency. cBased on the nominal operating characteristics listed in Table 1. 890


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Figure 1. Schematic of the Whispergen DC micro-CHP system.

engines, with a particular emphasis on emission reductions.23−26 The use of ethanol is generally accompanied with reductions in both particulate and CO emissions. However, effects on UHC and NOx emissions can vary greatly depending on the mode of combustion, the type of engine, and the specific operating conditions. Ethanol was chosen as a viable replacement for diesel in this study because it is generally more widely available than biodiesel. 1.4. Objectives. The focus of the current research was to compare the performance of a residential microcogeneration system based on a Stirling engine and fueled by diesel and ethanol. This was accomplished by performing energy analysis, as well as particulate, unburned hydrocarbon, carbon monoxide, nitric oxide, nitrogen dioxide, formaldehyde, acetaldehyde, and methane emission measurements. This experimental study was conducted on the Whispergen DC microcombined heat and power system, originally designed for use with diesel fuel. Engine operation with ethanol required the development of a new fuel delivery system to compensate for ethanol’s much lower heating value. In addition, the engine was optimized for use with ethanol through modifications to various engine parameters including initial fuel flow, maximum fuel flow, and glow plug duration. Lastly, parametric studies on primary engine set points, including coolant temperature and exhaust temperature, were conducted to further understand their effect on engine behavior with respect to efficiency and emissions.

Table 3. Specifications of the Whispergen DC Micro-CHP System feature

specificationa

prime mover engine configuration working fluid heat output power output fuel fuel consumption exhaust temperature size

four-cylinder Stirling engine double-acting alpha-type with wobble yoke mechanism nitrogen pressurized to 2.8 MPa 5.5 kW (nominal) 0.8 kW (nominal) No. 2 diesel 1 L/hour (maximum) 80 °C (nominal) 450 (width) × 500 (depth) × 650 (height) mm3

a

Taken from literature source.4

along with radiation from the flame, preheats the air to approximately 500 °C. The burner features continuous premixed combustion with a single swirl evaporator, such that the fuel is not injected as a spray but rather enters an evaporator where it is heated, vaporized, and then mixed with air prior to combustion. Ignition is achieved using a glow plug, an element used to preheat the burner during the cold start of the engine. The Stirling engine has a four-cylinder alpha-type doubleacting configuration with nitrogen pressurized to 2.8 MPa as the working fluid. The pistons are made of alloy steel and are sealed using PTFE lip seals backed with O-rings. The engine’s hot-end heat exchangers are made of high-temperature stainless steel for corrosion resistance, while the cold-end heat exchangers are made of copper for high heat transfer rates. Also, the engine’s internal regenerators consist of a stainless steel mesh.4 The engine’s cylinders are repeatedly heated and cooled to produce expansion and contraction of the working fluid, resulting in piston motion. Preliminary testing of this engine has revealed a shaft speed of 1200−1500 rpm.27 A wobble yoke mechanism uses the linear motion to rotate the alternator, thus producing AC electricity. This mechanism was chosen since it produces very low piston side loads, incorporates prelubricated single degree of freedom bearings, and is easy to manufacture. The AC electricity produced by the alternator is converted to DC through the use of a rectifier.4,9 Next, the electrical output is stored in a 216 Ah Interstate 12 V DC deep cycle battery, which is then partially converted to 120 V of AC power using a 1500 W Xantrex inverter to power a 500 W portable work light. Thus, the work light acts as the electrical load during engine operation. A Seakamp copper multiple-pass shell and tube heat exchanger with a

2. EXPERIMENTAL METHODOLOGY 2.1. The Whispergen DC Microcombined Heat and Power (Micro-CHP) System. This performance study was conducted on a diesel fuel-fired Whispergen DC micro-CHP system. The system was designed by Whisper Tech Limited, a New Zealand firm that has developed Stirling engine-based microcogeneration systems for use in small-scale applications for over ten years. The system, shown in Figure 1, incorporates a burner, a Stirling engine, an alternator, a rectifier, an electronics enclosure (controller), a shell and tube heat exchanger, and an exhaust heat exchanger. Aspects of data acquisition are also shown, including the placement of a pressure gauge, thermocouples, and flowmeters. The specifications of the system are detailed in Table 3. The burner is composed of a series of shells designed to transfer heat from the exhaust gases to the incoming air. This heat transfer, 891


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Figure 2. Engine test procedure flow diagram. 2.1.3. Engine Operating Conditions. The engine features two primary modes of operation, auto-charge, and heat management. In auto-charge mode, the engine automatically starts when the battery charge level falls below a preset level and automatically stops when the battery is fully charged. Conversely, operation in heat management mode entails a set coolant temperature being maintained by the engine regardless of battery state. Heat management mode was chosen as it offers a set coolant temperature and, thus, a fixed production of heat during steady state operation. Also, discharging the battery at a rate comparable to the engine’s production of electricity enables steady state power production. The power generation of the Whispergen system is based on three stages of charging: bulk-charging, absorption-charging, and floatcharging. Engine operation begins with bulk-charging, a stage that offers the highest level of power generation. Both absorption-charging and float-charging stages feature a high level of battery bank charge (above 80%) and thus demand a lower level of power production. Fuel and air flow are ramped down significantly in these stages, reducing the engine’s power and thermal outputs. Both of these stages are not optimal, as a clamp resistive heater is used to maintain the set coolant temperature by converting some of the generated electricity into heat. In order to evaluate its full-load performance, the engine was kept in the bulk-charging stage by discharging the battery to a charge level of approximately 50% prior to an engine test. All engine tests in this study utilized heat management mode with bulk-charging for a test period of one hour. This duration was chosen to ensure ease of comparability with the results presented in the previous work of Aliabadi et al.10 In addition, the engine was operated with an exhaust temperature set point of 480 °C and an exhaust oxygen sensor set point between 8 and 9%. Figure 2 shows the engine test procedures used for each fuel. The operating stages of a diesel engine test consist of 12.5 min of preheating, 12.5 min of running up, 21.5 min of running, 6 min of running down, and 7.5 min of cooling off. In the preheating stage, a glow plug is used to preheat the burner. The glow plug remains on until the 12.5 min mark, even though the fuel flow is initiated at the six minute mark. For ethanol, glow plug operation was extended by 2.5 min to facilitate the preheating stage and prevent blow-out of the flame. As a result, ethanol’s preheating stage overlaps with the running up stage, a period marked with the start of power generation and its subsequent rise. The engine is considered to be in the running stage when the power generation climbs to over 1 kW around the 25 min mark. Following the running stage, the engine is then shut down at the 46.5 min mark. Electricity is still generated for approximately six minutes, as the combustion chamber is still relatively hot. Finally, the alternator stops generating electricity, and the air blower forces air through the system to cool the engine. 2.2. Fuel Properties. The specific fuels used in this study were No. 2 diesel fuel and pure ethanol. Table 4 outlines significant differences

3.125 in. diameter and a 20 in. length is used to load the thermal output, as heat is removed from the system by running cold water through the shell. The engine coolant passes through the tubes in counterflow after it extracts heat from both the engine block and the exhaust heat exchanger. In addition, a clamp element heater is built into the coolant circuit to provide additional heating in certain modes of operation. The exhaust system consists mainly of an exhaust heat exchanger, a manometer, a stack fan, a variety of heaters, and additional parts related to sampling. To monitor the exhaust gas pressure, a Dwyer Instruments inclined-scale manometer was installed near the exhaust heat exchanger. Throughout an engine test, a stack fan is used to reduce back pressure. The flow rate of the stack fan is varied using a variable voltage output transformer to ensure that the pressure is always slightly below atmospheric, thereby preventing potential exhaust leaks. It should be noted that the system schematic highlights components that are crucial for engine operation. The engine’s two primary sensors, the exhaust oxygen sensor and the exhaust temperature sensor, are located between the burner and the exhaust heat exchanger. The exhaust oxygen sensor is responsible for maintaining a fixed fuel−air equivalence ratio (0.55−0.60), while the exhaust temperature sensor maintains the exhaust temperature set point (480 °C). 2.1.1. Development of the Fuel Delivery System. The Stirling engine’s original fuel system is based on a 12 V fuel pulse pump capable of delivering a diesel fuel flow rate of 19 mL/min. Since ethanol has a much lower energy content compared to diesel fuel, an increase in the fuel flow rate of ethanol was required to provide an equivalent amount of fuel energy. As a result, the Stirling engine’s fuel pump was replaced with a Cole Parmer Masterflex peristaltic pump with a computer-compatible drive. 2.1.2. Data Acquisition. The engine is equipped with a multitude of sensors, outputs of which include glow plug operation, exhaust temperature (prior to the heat exchanger), coolant temperature, burner temperature (interface between the burner and engine block), power output, and many more. Micromon V1.0 software was employed to observe the performance of the Whispergen system, log these engine parameters, and adjust key engine parameters, including exhaust temperature set point, oxygen sensor set point, initial fuel flow, and glow plug duration. The air flow rate of the engine was measured using an Eldridge Products air mass flowmeter with a reported accuracy of ±5−8 L/min under standard operating conditions. The fuel flow rate was determined through calibration of the peristaltic pump before and after each test, resulting in diesel and ethanol fuel flow rate errors of ±0.5% and ±0.6%, respectively. In addition, system temperatures were measured using Omega sheathed J-type thermocouples with grounded junctions and an accuracy of ±0.3 °C. Furthermore, the water flow rate was monitored using an Omega rotameter, accurate to within 6%. 892


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Table 4. Comparison of Diesel and Ethanol Fuel Properties

a

property

diesela

ethanola

lower heating value (MJ/kg) higher heating value (MJ/kg) autoignition temperature (°C) heat of vaporization (kJ/kg) 90% distillation/boiling temperature (°C) C−H−O composition (mass %)

42.8 45.6 204−260 225−600 282−338 86−14−0

26.9 29.7 365−425 837 78−79 52−13−35

Taken from literature sources.28,29

in fuel properties, including those based on energy content (lower and higher heating values), ignition characteristics (autoignition temperature and heat of vaporization), fuel volatility (90% distillation/boiling temperature), and chemical composition (carbon, hydrogen and oxygen content). Some of the diesel fuel properties were obtained from fuel certification tests conducted by the Alberta Research Council and established by the American Society for Testing and Materials, and the remainder are reported general properties. It can readily be seen that ethanol has much lower heating values, indicating its fuel consumption must be considerably larger than that of diesel to provide the same amount of power output. Regarding ignition characteristics, it is well-established that ethanol’s high autoignition temperature and heat of vaporization pose difficulties in achieving flame stability during the cold start of an engine. Fortunately, ethanol is a fully distillable fuel with a highly oxygenated chemical structure, generally allowing for more complete combustion and a substantial reduction in particulate emissions and certain exhaust species emissions. 2.3. Energy Efficiency. Energy efficiency is defined as the fraction of a fuel’s input energy that can be recovered in the form of heat and power. It can be obtained from eq 1, where Ẇ net is the net power generated, Q̇ R is the heat recovery rate, ṁ R is the fuel’s mass flow rate, and hF is the fuel’s specific enthalpy. The net power is calculated from the product of voltage and current produced by the electrical generator. The remaining heat energy is used to warm up water flowing through the shell-and-tube heat exchanger (see Figure 1). This energy is extracted by coolant which flows through the engine block and the exhaust heat exchanger. The high total efficiencies that have been calculated are justified by the relatively low temperature of 80 °C at which exhaust gas exits the system, indicating good overall utilization of energy. Depending on the application, cogeneration systems can be used to meet electrical demands, thermal demands, or both. In accordance, expressions for both power efficiency and thermal efficiency are included in this equation

Figure 3. Schematic of the particulate sampling system. mass of the particulates collected on the filters was measured by the Ohaus Analytical Plus microbalance. Appropriate measures were taken to ensure that sampling was consistent and reproducible. Prior to an engine test, the filter was weighed a total of three times. To get an accurate particle emissions rate during steady state operation, the particulates were collected between the 30 and 45 min marks. In this time interval, air flow, fuel flow, power output, and most engine parameters change only slightly. Following the test, the filter was placed in a crucible and allowed to dry for one day to remove water and volatile compounds present on the filter. The filter was again weighed three times to determine the mass of the particulate sample. Lastly, appropriate data were utilized to determine particulate concentration, emissions rate and specific emission based on total energy output. Finally, a statistical analysis was incorporated to examine the accuracy of the measurements. 2.5. The Flame Ionization Detector. A flame ionization detector (FID) is an extremely sensitive instrument used to measure total unburned hydrocarbon (UHC) emissions. The FID is an effective hydrocarbon counter by virtue of its ability to burn UHCs present in the exhaust sample in a small hydrogen-air flame. This produces ions proportional to the number of carbon atoms in hydrocarbon species found in the exhaust.28 In this study, total unburned hydrocarbon emissions were measured using the California Analytical Instruments model 600 HFID heated total hydrocarbon analyzer. The FID was calibrated using moisture-free air and 104 ppm (ppm) methane in nitrogen. A detection range of 0−300 ppm with an accuracy of ±3 ppm and a lower detection limit of 0.1 ppm was achieved through instrumental extrapolation. UHC emissions were reported as a mole fraction in ppm of methane. The exhaust sampling system of the FID is presented in Figure 4. The flame ionization detector’s built-in pump draws a fraction of the exhaust gas, approximately 1.5 L/min, through a heated filter, a heated transfer line, and a series of heated components before entering the FID. A stainless steel three-way valve was used to set the exhaust line for either UHC emission testing by the flame ionization detector or species emission testing by the Fourier transform infrared spectrometer. A Unique Heated Products heated filter and heated transfer line were used with a variety of heat tapes to keep the exhaust system at a relatively high temperature, generally at or above 190 °C, to avoid the condensation of higher boiling point range hydrocarbons. However, since the Stirling engine exhaust exit temperature was measured at approximately 70 °C, some condensation was likely. Fortunately, the production of heavier, higher boiling point range hydrocarbons should be limited, as the engine features continuous premixed combustion with a considerable amount of air preheat, approximately 500 °C. Nevertheless, relatively high sampling temperatures were applied to

̇ + Q̇ Wnet R ṁ FhF Q̇ R Ẇ = net + ṁ FhF ṁ FhF

ηenergy =

= ηpower + ηthermal

(1)

2.4. Particulate Matter Collection. Particulate matter is mainly composed of combustion generated carbonaceous material. It causes a variety of health problems including asthma, bronchitis, and other respiratory diseases.30 The particulate sampling system employed in this study was originally developed by Aliabadi.31 Certain modifications were made to the system, including a redesign of the sampling probe and changes to various components within the system. A schematic of the particulate sampling system is presented in Figure 3. A flowmeter with a control valve was used in conjunction with a vacuum pump to draw the exhaust through a tapered and machined stainless steel probe, a heated sample line, and a particulate filter. Direct isokinetic sampling of the exhaust was made at a fixed flow rate and temperature (56.6 L/min and 55−60 °C, respectively). Particulates were collected on Pall Corporation EMFAB filters which provide 99.9% retention for particulate sizes larger than 0.3 μm. The 893


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Figure 5. Schematic of the FTIR sampling system. pressure of 86.325 kPa, and a flow rate of 10.3 L/min. Each spectrum was acquired with 24 successive scans over approximately one minute at an optical resolution of 1 cm−1 with a spectral range of 1100−4000 cm−1. During the collection of a single spectrum, the gas cell was refilled approximately 51 times. Thus, each sample was an averaged representation of the exhaust. 2.7. Engine Operation. Although most engine parameters reached a steady state value relatively quickly, the engine’s thermal output and certain system temperatures took a long time to achieve equilibrium. Therefore, values at steady state were defined as the average value in the last minute of engine operation, between the 45.5 and 46.5 min marks. Table 6 lists steady state engine parameters

Figure 4. Schematic of the FID exhaust sampling system. prevent further condensation. The sample temperature at the inlet of the FID was between 180 and 182 °C throughout all tests. 2.6. The Fourier Transform Infrared (FTIR) Spectrometer. The Thermo Scientific Nicolet 380 FTIR spectrometer was used to measure carbon dioxide (CO2), water (H2O), carbon monoxide (CO), methane (CH4), nitric oxide (NO), nitrogen dioxide (NO2), formaldehyde (CH2O), and acetaldehyde (C2H4O) emissions. The analyzer operates by subjecting a sample to infrared radiation at a multitude of frequencies to determine the specific frequencies at which the sample absorbs radiation and the associated intensities. By developing an absorption-frequency curve, referred to as an infrared spectrum, one can subsequently identify which species are present in the exhaust and in what concentrations. The spectrometer is equipped with a potassium bromide beam splitter and a thermoelectrically cooled high performance deuterated triglycine sulfate detector. In addition, the system incorporates a heated Specac gas cell with a path length of 2 m and a cell volume of 0.19 L. A partial least-squares (PLS) calibration model was developed that incorporated a reasonable number of standards with an emphasis on complex multicomponent mixtures. Gas standards were generated using single gas cylinders (CO2, CO, CH4, NO, and NO2), permeation tubes (CH2O and C2H4O), and a gastight syringe (H2O). Table 5

Table 6. Engine Parameters with Associated Standard Deviations

analysis range (cm−1)

unit

lower limit

upper limit

RMSE

CO2 H2O CO CH4 NO NO2 CH2O C2H4O

2242.00−2209.00 3935.00−3926.50 2185.00−2164.00 3018.50−3004.50 1908.00−1881.00 1610.00−1584.00 2888.00−2790.50 2751.00−2655.00

% % ppm ppm ppm ppm ppm ppm

2.00 0.50 10.00 10.00 20.00 10.00 10.00 30.00

15.00 15.00 250.00 250.00 250.00 250.00 150.00 150.00

0.1 0.1 5.6 3.1 5.2 2.5 1.7 4.7

diesel

ethanol

fuel flow rate (mL/min) air flow rate (L/min) fuel−air equivalence ratio coolant temperature (°C) burner temperature (°C) hot exhaust temperature (°C) cooled exhaust temperature (°C)

16.0 (0.2) 279 (3) 0.58 (0.01) 65.3 (0.4) 429.9 (1.0) 479.1 (0.8) 71.6 (0.1)

26.9 (0.2) 283 (3) 0.56 (0.01) 62.2 (1.9) 425.7 (0.3) 477.0 (1.6) 73.0 (0.1)

averaged over three separate engine tests for each fuel, along with their respective sample standard deviations. The fuel consumption of ethanol was much greater at 26.9 mL/min compared to diesel fuel (16.0 mL/min). This was not unexpected, as the lower energy content of ethanol required an increase in fuel consumption by a factor of 1.7 on a volumetric basis to equal the fuel energy input of diesel. Otherwise, most engine parameters were similar for both fuels, with minor differences in certain system temperatures. For instance, variations in coolant temperature were primarily associated with the day-to-day variability of the building water’s flow rate (5.7−6.4 L/min) and inlet temperature (10−21 °C). Ambient temperatures were between 20 and 25 °C throughout engine testing.

Table 5. Summary of the Partial Least Squares Calibration Model species

parameter

3. RESULTS AND DISCUSSION

provides a summary of the PLS calibration model, listing the upper and lower detection limits, the root mean squared error (RMSE), and the spectrometer’s analysis range for each species. The lower detection limit of each species represents the concentration that would produce a signal-to-noise ratio of 3. A schematic of the FTIR sampling system is presented in Figure 5. It features a stainless steel manifold mainly used for exhaust sampling and a brass manifold strictly used for calibration. The entire system was partially insulated in addition to being heated with a variety of heat tapes. During an engine test, the exhaust was sampled and delivered to the spectrometer at a temperature of 115−120 °C, an absolute

3.1. Energy Efficiencies. Figure 6 illustrates the transient nature of the system’s power, thermal and energy efficiencies when powered by diesel and ethanol. A considerable amount of time was needed for system efficiencies to reach their steady state values. Since water present in the exhaust was in the vapor phase, efficiencies were computed using the more appropriate lower heating value (LHV). It must be stressed that actual power efficiencies are slightly lower than those reported here, since the power output logged by the software does not 894


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that engine operation with ethanol resulted in a reduction in burner temperature of approximately 4−5 °C. This is likely due to the fact that ethanol flames are less luminous and have a lower flame temperature compared to diesel. This reduces heat transfer from the combustion chamber to the hot-end heat exchangers and working fluid, leading to a slight reduction in power efficiency. However, the excess energy not transferred to the engine block in the case of ethanol is eventually recovered by the exhaust heat exchanger, which accounts for the higher thermal efficiency and similar overall efficiency compared to diesel operation. Upon closer inspection, the greater thermal energy recovery of the ethanol-powered system was partly due to slight differences in cooling. As previously shown in Table 6, a reduction in coolant temperature of approximately 3 °C was observed with ethanol, likely a result of a slightly higher building water flow rate. As efficiencies rely heavily on fuel flow rate, power output ,and thermal output, all parameters that fluctuate slightly during steady state operation, multiple engine tests were conducted to verify the accuracy of these measurements. Examining energy data from Table 7 reveals that while both power output and power efficiency were quite reproducible, thermal output, thermal efficiency, and energy efficiency had moderate levels of variance. The reproducibility of the power efficiency is attributable to the fact that it depends on the fuel flow rate and the engine’s power output, both of which were verified to be highly accurate readings. The computation of the thermal efficiency depended on the fuel flow rate, the water flow rate, the inlet water temperature, and the outlet water temperature. Variations in thermal energy were primarily associated with the fluctuating nature of the building water’s flow rate and inlet temperature and the inability to detect such minor changes. In summary, the power, thermal, and energy efficiencies of the system run on diesel and ethanol were similar on a LHV basis. 3.2. Particulate Emissions. The particulate emissions data for both diesel and ethanol are summarized in Table 8. Run on

Figure 6. Lower heating value efficiencies of the system run on diesel and ethanol.

account for the power consumption of ancillary components. Fortunately, efficiency comparisons are still accurate because both fuels have similar levels of power consumption during an engine test, with the exception of the preheating stage. The steady state power, thermal, and energy efficiencies of the system run on diesel and ethanol are presented in Table 7 on the basis of the LHV. Average energy data over three separate engine tests were used for each fuel to ensure a high level of reproducibility. The standard deviation for the total efficiency was obtained by adding the standard deviations of the 2 power and thermal efficiencies in quadrature (σenergy = [σpower + 1/2 2 σthermal ] ). The fraction of fuel energy that was not recovered was mainly due to radiation and convection losses within the burner and engine, as well as the loss of low temperature heat within the exhaust. The low power efficiencies were primarily due to the choice of working fluid and its mean cycle pressures. Nitrogen’s poor heat transfer and fluid friction properties make it an inferior working fluid compared to either hydrogen or helium in regard to power density and efficiency.27 Also, mean cycle pressures in the range of 10−20 MPa are required to achieve high power efficiencies, whereas the working fluid of this engine was pressurized to only 2.8 MPa.32 A similar study by Aliabadi et al.10 revealed a thermal output of 7.4 kW and a LHV energy efficiency of 90.5% when powered by diesel fuel, values larger than those obtained in this study. This is likely a consequence of their measurements being recorded during the absorption-charging stage in which electricity from the battery is used to generate heat. Operation in this regime is known to result in an overestimation of the thermal and energy efficiencies. Engine operation with ethanol resulted in lower power efficiency and higher thermal efficiency. This slight difference can be easily explained by the engine’s method of power production. For this particular Stirling engine, power control is managed by utilizing an exhaust gas temperature set point which indirectly regulates the temperature at the hot-end heat exchangers in the engine block. The closest measure of the temperature at the heat exchangers is the temperature at the interface between the burner and engine block, otherwise known as the burner temperature. It is evident from Table 6

Table 8. Particulate Emissions Data with Associated Standard Deviations fuel

sample mass (mg)

concentration (mg/m3)

emissions rate (mg/h)

specific emission (mg/kWh)

diesel ethanol

0.16 (0.04) 0.15 (0.03)

0.19 (0.05) 0.18 (0.04)

3.16 (0.80) 3.00 (0.60)

0.42 (0.11) 0.40 (0.08)

diesel, the particulate concentration, emissions rate, and its specific emission based on total energy output were 0.19 mg/ m3, 3.16 mg/h, and 0.42 mg/kWh, respectively. Equivalent particulate emissions data for ethanol were nearly identical at 0.18 mg/m3, 3.00 mg/h, and 0.40 mg/kWh, respectively. Emissions were extremely low, as the combustor features a continuous premixed flame that facilitates the complete burnout of already evaporated fuel. Interestingly, diesel particulate emissions of the Whispergen system measured by Aliabadi et al.10 were proportionally higher with a specific emission of 2.33 mg/kWh, as measurements were made over the entire engine test. In contrast, particulate emission testing

Table 7. Energy Outputs and LHV Efficiencies with Associated Standard Deviations fuel

power output (W)

thermal output (W)

power efficiency (%)

thermal efficiency (%)

energy efficiency (%)

diesel ethanol

1142 (7) 1117 (8)

6934 (51) 7024 (19)

12.1 (0.1) 11.8 (0.1)

73.3 (0.8) 73.9 (0.5)

85.4 (0.8) 85.7 (0.5)

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Figure 8 is a graphical representation of the total UHC emissions for ethanol. The system’s operating stages are

in this study was conducted only during the running stage when combustion was complete and highly stable. Visual inspection of both particulate filters revealed no evidence of any particulate matter. Although a reduction in particulate emissions was expected through the use of ethanol due to its highly oxygenated nature, the mode of combustion resulted in negligible particulate emissions for both fuels. 3.3. Unburned Hydrocarbon Emissions. Two unburned hydrocarbon (UHC) emission tests were conducted for each fuel. The initial test will be used to explain emissions behavior, while the additional test is included to assess the reproducibility of the results. Figure 7 illustrates the total diesel UHC

Figure 8. Ethanol total unburned hydrocarbon emissions.

equivalent to diesel with one exception; glow plug operation was extended by 2.5 min to facilitate the preheating stage and combat ethanol’s difficulties with ignition. During the early part of the preheating stage (0−6 min), the UHC emissions were below the instrument’s detection limit of 0.1 ppm. Upon fuel injection (6 min mark), they reached the maximum saturated value of 300 ppm and persisted at this level 4 min longer than diesel. The UHC emissions fell to 63 ppm by the end of the preheating stage (6−15 min) and rose to 230 ppm as the glow plug was turned off (15 min mark). Interestingly, emissions at the end of the running up stage (12.5−25 min) and for the duration of the running stage (25−46.5 min) were below the instrument’s detection limit, which is similar to diesel fuel. The UHC emissions peaked at the maximum saturated value when the flame was extinguished, but emissions dropped below the instrument’s detection limit by the end of the running down stage (46.5−52.5 min) and for the duration of the cooling off stage (52.5−60 min). These results indicate that the UHC emissions for ethanol were much higher than for diesel up to the end of the running up stage (25 min mark). This is attributed to ethanol’s much higher autoignition temperature, as well as its higher heat of vaporization, both of which make the flame much less stable during the cold start process. However, the emissions during steady state operation fell under the instrument’s detection limit for both fuels. It is also of note that UHC emissions for diesel were found to be similar to those measured by Aliabadi et al.,10 though the steady state emissions in that study did not fall below the instrument’s detection limit. Instead, they measured a value of 7 ppm, which likely occurred as a result of unburned hydrocarbons present in the exhaust sample line. Here, the sample lines were heated and flushed prior to every engine test, thus achieving an UHC emission level under the instrument’s detection limit during steady state operation. The additional UHC emission test conducted for each fuel showed that while trends were particularly consistent, values differed slightly throughout the engine test. This can be explained by noting that the flame ionization detector displays a continuous reading for UHC emission levels. Since the data were logged with a sampling frequency of one minute, it was

Figure 7. Diesel total unburned hydrocarbon emissions.

emissions for a one hour test period with a sampling frequency of one minute. As previously outlined in Figure 2, the test is divided into five stages: preheating, running up, running, running down, and cooling off. In the preheating stage (0−12.5 min), the UHC emissions first reached a peak level of 32 ppm (5 min mark), as a glow plug was used to heat the evaporator. Emissions were likely due to the outgassing of fuel from the fuel line and the incomplete burning of fuel deposits near the heat source. Upon fuel injection (6 min mark), the UHC emissions reached the maximum saturated value of 300 ppm but quickly dropped to 37 ppm, coincident with the flame beginning to stabilize with the increase in fuel flow during the preheating stage. When the glow plug was turned off (12.5 min mark), the UHC emissions rose to 111 ppm, and the flame began to sustain itself by heat radiation from the internal surfaces of the combustion chamber. This was immediately followed by the running up stage (12.5−25 min) which involved a gradual increase in the engine’s power output. By the end of this stage and during the entire running stage (25−46.5 min), the UHC emissions fell below the instrument’s detection limit of 0.1 ppm. Upon termination of the flame, emissions reached the maximum saturated value but fell to 12 ppm by the end of the running down stage, which is marked with a steady reduction in the engine’s power output. This was followed by a decrease in the cooling off stage (52.5−60 min) and a subsequent drop below the instrument’s detection limit by the end of the test. The presence of emissions after the flame extinction can be attributed to the outgassing of fuel and the incomplete burning of fuel deposits on combustion chamber surfaces. 896


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condensed considerably, resulting in variability with readings obtained during transient operation. Additionally, both the outgassing of fuel from the fuel line and the incomplete burning of fuel deposits within the combustion chamber created variation in UHC emission levels during the preheating, running down, and cooling off stages. 3.4. Exhaust Species Emissions. Figure 9 represents the various species emissions for diesel. During the early part of the

Figure 10. Ethanol exhaust species emissions.

negligibly small NO2/NO ratios for typical flame temperatures. Conversion of NO to NO2 was not observed in this study, though it has been demonstrated experimentally in exhaust sample lines.33 An additional species emission test was conducted for each fuel to acquire more accurate emissions data, as well as to assess the variability associated with emissions during steady state operation. Table 9 lists steady state CO and NO emissions

Figure 9. Diesel exhaust species emissions.

preheating stage (0−6 min), most pollutants were present below the spectrometer’s detection limit. At the start of fuel injection and during the rest of the preheating stage (6−12.5 min), most pollutants reached their maximum levels, with the exception of nitric oxide. The maximum pollutant levels were 250 ppm CO (saturated), 54 ppm CH4, 25 ppm CH2O, and 51 ppm C2H4O. During the running up and running stages (12.5− 46.5 min), all species emissions were below the spectrometer’s detection limit except those for nitric oxide and carbon monoxide. By the end of the running stage, emissions of CO and NO reached an average of 95 ppm and 80 ppm, respectively. Though there was a sharp increase in various species emissions upon flame termination, all species emissions dropped below the spectrometer’s detection limit by the end of the running down stage. Species emissions for ethanol are displayed in Figure 10. All pollutant species were present below the spectrometer’s detection limit during the early part of the preheating stage. At the start of fuel injection, most pollutants reached their maximum levels, with the exception of nitric oxide. The maximum pollutant levels were 250 ppm CO (saturated), 60 ppm CH4, 77 ppm CH2O, and 150 ppm C2H4O (saturated). Due to ethanol’s high autoignition temperature and heat of vaporization, most species emissions for ethanol were much higher than diesel up to the end of the running up stage (25 min mark). However, all species emissions during steady state operation fell below the spectrometer’s detection limit, with the exception of NO and CO. By the end of the running stage, CO and NO emissions reached an average of 65 ppm and 22 ppm, respectively. It is of note that behavior of the pollutants in the running down and cooling off stages is very similar to diesel. Nitrogen dioxide emissions were not included in the species emission plots, since they were below the spectrometer’s detection limit throughout the entire test for both fuels. This was reasonable, as chemical equilibrium considerations indicate

Table 9. CO and NO Emissions with Associated Standard Deviations fuel

CO emissions (ppm)

NO emissions (ppm)

diesel ethanol

95.3 (1.7) 67.1 (5.5)

83.8 (6.6) 23.3 (3.8)

averaged over two separate engine tests for each fuel. As emission levels reached steady state quicker than efficiency measurements, steady state values were defined as averages between the 40 and 46 min marks. Species emissions between engine tests varied by as much as 8%, which is primarily attributed to the variability associated with combustion. Issues with the calibration model could also play a role; while CO exhibits a nonlinear relationship between concentration and absorbance, NO has substantial interference with water. Nevertheless, emissions behavior was reasonably consistent and reproducible. During steady state operation, the CO and NO emissions for ethanol (67 and 23 ppm or, equivalently, 50 and 19 mg/m3) were much lower than diesel (95 and 84 ppm or, equivalently, 71 and 67 mg/m3). Lower CO emissions associated with ethanol combustion can be justified by observing the method of fuel/air mixing and the difference in the gas jet velocities of the fuels. As the fuel enters the evaporator, it is subsequently heated and vaporized. The vaporized fuel then leaves the evaporator and comes into contact with air for a relatively short period of time. This occurs prior to its entry into the flame, in a process of fuel/air mixing resembling that of a reacting crossflow. Since reacting crossflows are complex and involve the interaction of several phenomena, nonreacting crossflow correlations are typically used as a guideline to determine the degree of mixing and its effect on emissions. An important parameter that relates well to mixing is the momentum flux ratio (J), which is the ratio of the momentum of the jets to the 897


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Table 10. Energy Outputs and LHV Efficiencies for Coolant Temperature Tests with Associated Standard Deviations coolant temperature (°C)

water flow rate (L/min)

power output (W)

thermal output (W)




52.2 (0.2) 61.8 (0.1) 73.1 (0.1)

8.3 6.1 3.8

1118 (4) 1125 (6) 1107 (5)

7243 (19) 7021 (27) 6748 (29)

11.8 (0.1) 11.8 (0.1) 11.7 (0.1)

76.7 (0.7) 73.7 (0.5) 71.1 (0.6)

88.5 (0.7) 85.5 (0.5) 82.8 (0.6)

Table 10 summarizes engine performance as a function of steady state coolant temperature. While power output and efficiency did not vary with respect to coolant temperature, coolant temperature had a significant effect on the thermal output, thermal efficiency, and energy efficiency. Specifically, a drop in coolant temperature of approximately 10−11 °C was accompanied with a 2−3% increase in thermal efficiency and thus energy efficiency, as the Stirling follows the definition of a Carnot cycle. 3.6. Exhaust Temperature Study. A parametric study on exhaust temperature is critical, as engine operation is primarily based on the exhaust temperature set point. Section 3.1 discussed how power control of this Stirling engine is achieved by regulating the temperature at the hot-end heat exchangers within the engine block. In addition, the lower power output with ethanol operation was attributed to its lower adiabatic flame temperature and radiative power, which resulted in a lower burner temperature. Thus, through engine operation at higher exhaust temperatures, the power generation associated with ethanol can be made equivalent to that of diesel. Varying the exhaust temperature set point using the engine’s software was ineffective, since the engine is highly optimized and designed to run at a fixed operating condition. To circumvent this limitation, the exhaust temperature was varied by adjusting the oxygen sensor set point by a small amount. An engine test was conducted with a slight increase in the set point to run the engine with leaner mixtures, thus resulting in reduced combustion chamber temperatures. Alternatively, higher combustion temperatures were achieved by decreasing the set point to yield slightly richer operating conditions. A summary of the relevant steady state parameters for each engine test is provided in Table 11. As expected, an increase in

momentum of the crossflow (main flow). The correlation is illustrated in eq 2 where ρ and v are the density and velocity, respectively. It is evident that an increase in the gas jet velocity is accompanied with an increase in the momentum flux ratio, and, thus, a higher degree of mixing and a reduction in CO emissions is achieved for underpenetrating jets.34 This is accomplished with ethanol as its gas jet velocity was determined to be six times that of diesel. In addition, the oxygenated nature of ethanol could also result in more complete combustion, with potential reductions in emissions

J=

ρjetvjet 2 ρmain flow vmain flow 2

(2)

Nitric oxide emissions were observed to increase throughout the engine tests as the burner temperature rose, demonstrating that the dominant mechanism for the formation of nitrogen oxides (NOx) was thermal NOx. Therefore, lower NO emissions for ethanol can be attributed to its lower flame temperature. Through the use of chemical equilibrium software, such as STANJAN, the flame temperature of ethanol is calculated to typically be 75−150 °C lower than that of diesel. Unfortunately, it is extremely difficult to get an accurate representation of the flame temperature for each fuel as the Stirling engine features very lean mixtures, a high degree of air and fuel preheat and considerable heat transfer from the combustion chamber to the engine block. The steady state emissions reported by Aliabadi et al.10 significantly differed from the values shown above. Powered by diesel, the engine produced 151 ppm NO, 25 ppm NO2, and 27 ppm CO during steady state operation. However, the partial least-squares calibration model developed in that study was based on pure compounds and thus did not properly account for the spectral interference between components within an exhaust sample. To account for this interference, the calibration model used here was based on a considerable amount of multicomponent standards, consisting of nearly all components present in the calibration. 3.5. Coolant Temperature Study. A parametric study was performed to examine the effect of coolant temperature on engine performance. This was accomplished by varying the flow rate of water that passes through the shell and tube heat exchanger in counterflow with the coolant. Two separate engine tests were conducted; one was operated with the coolant temperature higher than normal, the other lower. The engine was powered exclusively by ethanol, and, thus, the two tests were compared to a typical engine test operated with ethanol. Other than coolant temperature, the engine parameters for all tests were equivalent. This included the exhaust temperature set point, fuel flow, and air flow. As combustion was not affected by variations in coolant temperature, emission tests were not used in this study. Engine performance was compared by evaluating the energy outputs and LHV energy efficiencies during steady state operation.

Table 11. Engine Parameters for Exhaust Temperature Tests with Associated Standard Deviations exhaust temperature (°C)

fuel−air equivalence ratio

fuel flow rate (mL/min)

burner temperature (°C)

460.3 (0.2) 478.3 (0.3) 493.2 (0.2)

0.54 (0.01) 0.56 (0.01) 0.59 (0.01)

25.9 (0.2) 26.9 (0.1) 27.8 (0.2)

410.2 (0.1) 425.8 (0.3) 441.8 (0.3)

the fuel−air equivalence ratio and overall fuel flow was required to achieve a higher exhaust temperature. This was accompanied by an increase in burner temperature, which is expected to correlate with the power efficiency of the engine. Table 12 lists steady state energy outputs and LHV efficiencies for each engine test. Increasing the exhaust temperature by 15−18 °C was accompanied by a 1.3−1.4% increase in thermal efficiency and a slight increase in power efficiency. As previously mentioned, this increase in thermal efficiency can be attributed to the greater temperature differential across the exhaust heat exchanger. The slight increase in power efficiency was due to the greater temperature difference between the hot and cold ends of the Stirling engine.27 898


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Table 12. Energy Outputs and LHV Efficiencies for Exhaust Temperature Tests with Associated Standard Deviations exhaust temperature (°C)

power output (W)

thermal output (W)




460.3 (0.2) 478.3 (0.3) 493.2 (0.2)

1053 (6) 1125 (6) 1182 (5)

6641 (21) 7021 (27) 7394 (21)

11.5 (0.1) 11.8 (0.1) 12.0 (0.1)

72.4 (0.7) 73.7 (0.5) 75.1 (0.6)

83.9 (0.7) 85.5 (0.5) 87.1 (0.6)

higher than diesel up to the 25 min mark. However, all species emissions were below the spectrometer’s detection limit during steady state operation with the exception of nitric oxide (NO) and carbon monoxide (CO). Emissions of CO and NO for ethanol (50 and 19 mg/m3, respectively) were found to be much lower than that of diesel (71 and 67 mg/m3, respectively). These lower NO emission levels were attributed to ethanol’s lower flame temperature, whereas reductions in CO emissions could be due to a higher degree of fuel/air mixing with ethanol, as a result of higher gas jet velocities of the fuel exiting the orifices of the evaporator. Lastly, through a series of engine tests powered by ethanol, parametric studies on coolant temperature and exhaust temperature were conducted, revealing a decrease in coolant temperature was accompanied with a substantial increase in thermal efficiency. In contrast, an increase in exhaust temperature resulted in a slight increase in power efficiency and a moderate increase in thermal efficiency, with little to no effect on species emissions. Engine performance with ethanol resulted in considerable emission reductions during steady state operation, with energy efficiencies comparable to those of diesel. However, this occurs at the expense of fuel consumption, which is increased by a factor of 1.7 on a volumetric basis to compensate for ethanol’s lower energy content.

Since the fuel energy input and equivalence ratio for each engine test varied, it was necessary to conduct species emission tests. Table 13 outlines the steady state emissions for each Table 13. CO and NO Emissions for Exhaust Temperature Tests with Associated Standard Deviations exhaust temperature (°C)

CO emissions (ppm)

NO emissions (ppm)

460.3 (0.2) 478.3 (0.3) 493.2 (0.2)

62.5 (3.2) 64.5 (5.8) 66.1 (3.6)

17.6 (2.4) 22.1 (3.6) 23.2 (5.1)

engine test. Only carbon monoxide and nitric oxide emissions are presented, as there were no other steady state emissions. While a slight rise in CO and NO emissions was observed alongside the increase in exhaust temperature, emission levels overlap when considering measurement errors, as well as instrumental errors. In addition, NO emissions were found to be at or below the spectrometer’s detection limit of 20 ppm. Thus, an increase in exhaust temperature was accompanied with an improvement in both thermal and power efficiencies with negligible effects on steady state emissions.

4. CONCLUSIONS This study comparatively examined the performance of the Whispergen DC microcombined heat and power system when fueled by diesel and ethanol. Based on the lower heating value, the engine achieved a power efficiency of 12.1%, a thermal efficiency of 73.3%, and a total efficiency of 85.4% when powered by diesel fuel. Corresponding efficiencies for ethanol were 11.8%, 73.9%, and 85.7%, respectively. Although the total energy efficiency for each fuel was similar, engine operation with ethanol resulted in lower power efficiency and higher thermal efficiency. The lower power efficiency associated with ethanol combustion was likely a result of reduced heat transfer from the combustion chamber to the working fluid, since an ethanol flame has lower adiabatic flame temperature and radiative power when compared to diesel. The slightly higher thermal efficiency was attributed to additional heat extraction at the exhaust exchanger, but this was partly a result of differences in cooling caused by the day-to-day variability of the building water’s flow rate and inlet temperature. Particulate emissions were negligibly small, with specific emissions of 0.42 mg/kWh and 0.40 mg/kWh for diesel and ethanol, respectively. Although ethanol’s highly oxygenated nature is typically conducive to particulate emission reductions during engine operation, the use of continuous premixed combustion resulted in minimal particulate emissions for both fuels. Unburned hydrocarbon emissions for ethanol were found to be much higher than diesel from the start of fuel flow to the end of the running up stage (25 min mark), as the flame was much less stable during the cold start process. It is encouraging that steady state emissions of unburned hydrocarbons for both fuels were below the instrument’s detection limit of 0.1 ppm. Emissions of carbon monoxide, nitric oxide, nitrogen dioxide, formaldehyde, acetaldehyde, and methane for ethanol were

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AUTHOR INFORMATION

Corresponding Author

*Phone: 416 580 3391. Fax: 416 978 7753. E-mail: thomson@ mie.utoronto.ca.

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ACKNOWLEDGMENTS The authors thank Eric Schutte of Whisper Tech New Zealand for his help with the Stirling engine, particularly with understanding its inner workings and debugging technical issues. Financial support for this project was provided by NSERC.

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NOMENCLATURE

Abbreviations and Acronyms

CHP = combined heat and power FID = flame ionization detector FTIR = Fourier transform infrared (spectrometer) IC = internal combustion LHV = lower heating value NOx = oxides of nitrogen, NO and NO2 PLS = partial least-squares ppm = parts per million (volume) RMSE = root mean squared error UHC = unburned hydrocarbon Variables

hF = specific enthalpy of fuel J = momentum flux ratio ṁ F = mass flow rate of fuel Q̇ R = heat recovery rate vjet = velocity of gas jet flow 899


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vmain flow = velocity of cross-flow Ẇ net = net power Greek Letters

ηenergy = energy efficiency ηpower = power efficiency ηthermal = thermal efficiency ρjet = density of gas jet flow ρmain flow = density of cross-flow σenergy = standard deviation for energy efficiency σpower = standard deviation for power efficiency σthermal = standard deviation for thermal efficiency

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