||ABSORPTION REFRIGERATION SYSTEM USING ENGINE EXHAUST GAS
SEMINAR REPORT Submitted by
JESIL JOHN KURIAN
Department Of Mechanical Engineering
College of Engineering, Thiruvananthapuram – 16
ABSORPTION REFRIGERATION SYSTEM USING ENGINE EXHAUST GAS.docx (Size: 888.08 KB / Downloads: 1096)
This work presents an experimental study of an ammonia–water absorption refrigeration system using the exhaust of an internal combustion engine as energy source. The exhaust gas energy availability and the impact of the absorption refrigeration system on engine performance, exhaust emissions, and power economy are evaluated. A production automotive engine was tested in a bench test dynamometer, with the absorption refrigeration system adapted to the exhaust pipe. The engine was tested for 25%, 50%, 75% and wide-open throttle valve. The refrigerator reached a steady state temperature between 4 and 13 degree centigrade about 3 hours after system start up, depending on engine throttle valve opening. The calculated exhaust gas energy availability suggests the cooling capacity can be highly improved for a dedicated system. Exhaust hydrocarbon emissions were higher when the refrigeration system was installed in the engine exhaust, but carbon monoxide emissions were reduced, while carbon dioxide concentration remained practically unaltered.
INDEX OF CONTENTS
Sl. No. Title Pg. No.
Section 1 Nomenclature 6
Section 2 List of Figures 6
Section 3 Introduction 9
Section 4 Objective 10
Section 5 Absorption refrigeration System 11
Section 6 Experimental Setup 16
Section 7 Results and Discussions 18
7.1 Time variation of refrigerator average temperature 18
7.2 Time variation of relative humidity inside the
7.3 Time variation of cooling capacity 21
7.4 Engine exhaust gas power availability 22
7.5 Time variation of refrigerator coefficient of
7.6 Influence of refrigerator on engine performance 24
7.7 Influence of refrigerator on carbon oxides
7.8 Influence of refrigerator on hydrocarbon emissions 26
Section 8 Conclusion 27
Section 9 References 28
: Instantaneous coefficient of performance
: Absorption system instantaneous cooling capacity
: Instantaneous heat transfer rate from the energy source (in this work engine exhaust gas is used) to the absorption refrigeration system
: Instantaneous heat transfer rate from the neighborhood to the cold
source through the refrigerator walls
: Final system internal energy
: Initial system internal energy
: Time Interval
: Radiation from the neighborhood
: Natural convection from the environmental air
: The instantaneous internal energy
: The water vapor mass
: Partial pressure of water vapor
: Dry air mass
: The heat transfer rate from the engine exhaust gas to the
absorption refrigeration system vapor generator
: The engine exhaust gas mass flow rate
2. LIST OF FIGURES
Figure No: Name Pg. No.
1 Schematic of the ammonia–water
absorption refrigeration system 11
2 Absorption refrigerator adapted
to engine exhaust system 16
3 Time variation of refrigerator average
4 Time variation of relative humidity
inside the refrigerator 20
5 Time variation of cooling capacity 21
6 Engine exhaust gas power availability 22
7 Time variation of refrigerator coefficient
of performance 23
8 Influence of refrigerator on engine power
9 Influence of refrigerator on specific fuel
10 Influence of refrigerator on carbon
dioxide emissions 25
11 Influence of refrigerator on carbon
monooxide emissions 25
12 Influence of refrigerator on hydrocarbon
Energy efficiency has been a major topic of discussions on natural resources preservation and costs reduction. Based on estimates of energy resources reduction at medium and long terms, it is vital to develop more efficient processes from energy and exergy standpoints. Environment preservation must also be considered through energy optimization studies. An important point to mention absorption refrigeration systems is the continuing substitution of chlorinated fluorocarbons (CFCs) by alternative refrigerants, according to the Montreal Protocol, signed in 1987 by 46 countries and revised in 1990 to protect the ozone layer.
Other motivating factors are the continuous optimization of the performance of internal combustion engines and the increasing utilization of air conditioning in vehicles, as it reaches the status of essential need for modern life. Internal combustion engines are potential energy sources for absorption refrigeration systems, as about one third of the energy availability in the combustion process is wasted through the exhaust gas. Thus, use of the exhaust gas in an absorption refrigeration system can increase the overall system efficiency.
This work has as an objective the study of the feasibility and potential of using the internal combustion engine exhaust gas as energy source for an absorption refrigeration system. For this purpose was performed an experimental study on a commercial 215-l refrigerator. The impact of the absorption refrigeration system on engine power output and exhaust emissions is analyzed, in order to know how this system influences the operation of an internal combustion engine.
The demand for fossil fuels is on the rise and the threats possessed by the pollutants cannot be neglected. And so is the requirement for energy efficient machines and this topic deals with the usage of ‘wasteful’ energy from vehicular exhaust emissions for refrigeration or air conditioning purpose.
Air conditioning is also becoming a necessity in our society. Considering this, usage of different methods like absorption refrigeration systems, adsorption systems, solar systems, can contribute to the overall efficiency of vehicles
The topic also deals with an experiment related to a vehicle integrated with absorption refrigeration system. Its results and the scopes are also discussed in the topic.
5. ABSORPTION REFRIGERATION SYSTEM
Fig. 1. Schematic of the ammonia–water absorption refrigeration system.
Fig. 1 shows a schematic of the basic ammonia–water absorption refrigeration cycle. High pressure ammonia vapor enters the condenser, where it transfers heat to the neighborhood. Liquid ammonia leaves the condenser and passes through an expansion valve, reaching the evaporator pressure. The refrigerant then enters the evaporator, where it receives heat from the cold source, turning into low pressure vapor. In the sequence, ammonia vapor enters the absorber, where a weak solution of water and low concentration ammonia absorbs the refrigerant and, at the same time, transfers heat to the neighborhood. The solution has now a high ammonia concentration, and is pumped to the vapor generator, where it receives heat from an external source. The ammonia in the solution then evaporates, separating from water and flowing to the condenser to start a new cycle. A weak water–ammonia solution leaves the vapor generator and enters the absorber to absorb ammonia vapor from the evaporator. A heat exchanger between the absorber and the vapor generator transfers heat from the weak solution leaving the vapor generator to the high ammonia concentration solution going into the vapor generator. That increases the cycle coefficient of performance. The absorption refrigeration system instantaneous coefficient of performance (COP) is given by:
where Q_ ref is the absorption system instantaneous cooling capacity (W) and Q_ exh is the instantaneous heat transfer rate from the energy source (in this work engine exhaust gas is used) to the absorption refrigeration system (W). From the First Law of Thermodynamics, the system cooling capacity is calculated as:
where Q_ wall is the instantaneous heat transfer rate from the neighborhood to the cold source through the refrigerator walls (W), Ui and Uf are the initial and final system internal energy (J) at the time interval Dt (s), respectively. Q_ wall was calculated considering radiation from the neighborhood (Q_ rad;ext) and natural convection from the environmental air (Q_ conv;ext) to the refrigerator external walls. This energy was conducted through the refrigerator walls and transferred by natural convection from the refrigerator inner walls to the ambient air inside the refrigerator (Q_ conv;int). Thus,
where kwall is the refrigerator wall thermal conductivity (W/m K), Awall is the refrigerator wall area frontal to conduction heat flux(m2), dwall is the refrigerator wall thickness (m), hext is the external convection coefficient (W/m2 K), Asur,ext is the refrigerator total external surface area (m2), Tsur,ext is the refrigerator external surface average temperature (K), T1,ext is the environmental air temperature (K), e is the refrigerator external surface emissivity, r is the Stefan–Boltzmann constant (5.67 _ 10_8 W/m2 K4), Tneigh is the neighborhood temperature (K), considered to be equal to the environmental air temperature, hint is the internal convection coefficient (W/m2 K), Asur,int is the refrigerator total internal surface area (m2), Tsur,int is the refrigerator internal surface average temperature (K), and T 1,int is the ambient air temperature inside the refrigerator (K). Both the environmental air temperature and the ambient air temperature inside the refrigerator were measured, while the surface temperatures were determined by an iterative process using Eqs. (3)–(6). The remaining parameters were calculated as shown next.
Heat transferred from the environmental air to the refrigerator external walls and from the refrigerator internal walls to the ambient air inside the refrigerator were calculated considering natural convection to vertical (side surfaces) and horizontal (top and bottom surfaces) plates. For the vertical surfaces the convection coefficients were calculated by
For the refrigerator external top surface, the convection coefficients was so evaluated
And, for the refrigerator internal top surface, the convection coefficient was thus evaluated
where kf is the fluid thermal conductivity (W/m K), L is the characteristic length of the flat plate (m), RaL is the Rayleigh number, and Pr is the Prandtl number of the ambient or environmental air. Heat transfer through the refrigerator bottom wall was neglected. The instantaneous internal energy inside the refrigerator was evaluated by:
where mair is the dry air mass (kg), uair is the dry air specific internal energy (kJ/kg), mv is the water vapor mass (kg), uv is the water vapor specific internal energy (kJ/kg), ml is the condensed water mass (kg), and ul is the condensed water specific internal energy (kJ/kg). Both dry air and water vapor were treated as ideal gases. Thus, the water vapor mass is calculated as follows:
where pv is the partial pressure of water vapor in the ambient air inside the refrigerator (Pa), " is the measured volume inside the refrigerator (m3), R is the universal gas constant (8314 J/kmol K), Mv is the water vapor molecular weight (kg/kmol), and T is the average ambient temperature inside the refrigerator (K). The water vapor partial pressure is calculated by:
where / is the measured relative humidity of the ambient air inside the refrigerator and ps is the water vapor saturation pressure at the measured temperature of the ambient air inside the refrigerator. The dry air mass inside the refrigerator is evaluated in the following way:
where Mair is the dry air molecular weight (kg/kmol) and p is the absolute pressure inside the refrigerator (Pa). The specific internal energies uair, uv, and ul were taken from thermodynamic tables as function of average ambient air temperature inside the refrigerator. The water vapor and condensed water specific internal energies were considered as approximately equal to that of the saturated vapor and saturated liquid at the same temperature, respectively. The heat transfer rate from the engine exhaust gas to the absorption refrigeration system vapor generator is given by:
where m_ exh is the engine exhaust gas mass flow rate, hexh,in is the engine exhaust gas specific enthalpy at the entrance of the vapor generator and hexh,out is the engine exhaust gas specific enthalpy at the exit of the vapor generator. Q_ ins is the heat transfer from the engine exhaust gas to the environment though the insulated wall of the vapor generator, calculated in a similar procedure as described by Eqs. (3)–(8). The engine exhaust gas mass flow rate was determined through the following relationship:
where qF is the fuel density (kg/m3), determined as a function of the temperature, _ 8F is the measured fuel volume flow rate (m3/s), and A/F is the air/fuel ratio, that is, the ratio of air mass flow rate and fuel flow rate. The exhaust gas enthalpy variation was calculated as follows:
Where yi, hin, and hout are the mass fraction, inlet enthalpy (kJ/kg) and outlet enthalpy (kJ/kg) of the exhaust gas components CO, CO2, HC, H2O, O2, and N2. The exhaust gas components concentration was measured in the experiments, while the enthalpy wastaken from thermodynamic tables based on the exhaust gas temperature.
6. EXPERIMENTAL SETUP
Fig. 2. Absorption refrigerator adapted to engine exhaust system.
A commercial 215-l refrigerator built on an absorption refrigeration system was studied using the engine exhaust gas as energy source (Fig. 2). Temperature and humidity inside the refrigerator were monitored through two Pt100 thermometers and a thermohygrometer. Measurement of the exhaust gas temperature was made through two K-type thermocouples installed in the refrigeration system heat exchanger inlet and outlet. A barometer and a liquid- in-bulb thermometer were used to measure ambient pressure and temperature. Ambient pressure was kept at 0.913 ± 0.007 bar, while ambient temperature was maintained at 300 ± 5 K.
A production 1.6-l, 8-valve, four-cylinder automotive engine with multipoint electronic fuel injection was used for the tests. The engine also featured compression ratio 9.5:1, 86.4 mm bore and 67.4 mm stroke. The engine was tested in a hydraulic dynamometer of maximum power 260 kW and maximum speed 6000 rev/min. The dynamometer was equipped with a load cell of measuring range up to 2224 N and uncertainty of ±0.4 N, and a magnetic speed sensor, which uncertainty was ±3 rev/min. Fuel consumption was measured by a turbine flow meter, of measuring range 0.038–100 l per minute and accuracy of 0.5% of the reading. To adapt the refrigerator to the engine the exhaust pipe configuration after the catalytic converter had to be modified, including removal of two plenum chambers and a flexible duct that directed the exhaust gas to an external chimney. These components were part of the standard laboratory setting, but could not be used when the refrigerator was installed.
The exhaust concentrations of carbon monoxide (CO), carbon dioxide (CO2), hydrocarbons (HC) and oxygen (O2) were measured by a non-dispersive infrared analyzer (NDIR). The equipment also allowed for determination of air/fuel ratio and mixture equivalence ratio (ratio between measured air/fuel ratio and stoichiometric air/fuel ratio). The uncertainties associated to the measurements were ±0.16% for CO, ±0.14% for CO2, ±6.8 ppm for HC, and ±0.01% for O2. A computer-based data acquisition system was used to monitor humidity and temperature inside the refrigerator, evaporator inlet and outlet temperatures, and fuel flow rate. Air humidity was measured with an uncertainty of ±0.06%. Cooling capacity and coefficient of performance were evaluated with uncertainties of ±1.84Wand ±0.0014, respectively. All uncertainties of measurements were evaluated according to ABNT/INMETRO  (similar to NIST TN 1297), for a confidence level of 95%.
When powering the refrigerator with the engine exhaust gas it was verified that, for engine speeds over 2000 rev/min, the temperature in the refrigerator was increased. At such condition there was excessive energy being transferred from the high-temperature exhaust gas to the refrigerant, not allowing its condensation in the condenser due to the elevated sensible heat to be removed. As a consequence, the refrigerant temperature in the evaporator was above that inside the refrigerator. Thus, it was decided to perform the tests at a fixed engine speed of 1500 rev/min to avoid increasing temperature inside the refrigerator.
7. RESULTS AND DISCUSSIONS
7.1.Time variation of refrigerator average temperature
Figs. 3–6 show the results obtained from the absorption refrigeration system using the engine exhaust gas as energy source for 25%, 50%, 75% and wide-open engine throttle valve. More than 800 data points were used to build each curve shown by the figures.
Fig. 3. Time variation of refrigerator average temperature.
Fig. 3 presents the refrigerator average internal temperature at those engine operational conditions. In the first 30–45 min it is observed that the average refrigerator temperature increases with time. This phenomenon is characteristic of the system start-up working regime when there is a refrigerant hot flow inside the refrigerator evaporator that would heat the refrigerator interior while the refrigeration system cooling capacity increases (see Fig. 5). After 30–45 min, depending on engine valve opening, the refrigeration system reaches the steady state condition and its interior starts cooling down, leading to the reduction of the refrigerator average temperature (Fig. 3). The temperature dropped faster inside the refrigerator as the throttle valve opening was wider. Attainment of the refrigerator steady state temperature was faster as the engine throttle valve was widened. Overall, the refrigeration average internal temperature took around 3–3.5 h to reach the steady state condition. The wider the valve opening, the lower the steady state average temperature attained inside the refrigerator. This internal temperature varied between 5 and 13 _C. The absorption refrigeration system was shut down after approximately 3 h, when the steady state temperature was attained. After system shut down, the internal average refrigerator temperature started to increase. A longer period was required to attain a thermal equilibrium with the external environment temperature. This period was larger for widened engine valve openings.
7.2. Time variation of relative humidity inside the refrigerator
Fig. 4. Time variation of relative humidity inside the refrigerator.
Fig. 4 shows that, as the averaged temperature inside the refrigerator was risen from start up (see Fig. 3), the relative humidity was also increased for all valve openings. A high slope reduction on the relative humidity inside the refrigerator was observed from the time the temperature started to fall (see Fig. 3), until reaching the refrigerator steady state condition. This trend is attributed to water vapor inside the refrigerator being condensed and, eventually, turning into ice with temperature decrease, thus reducing air humidity. The steady state relative humidity recorded was between 29%, for 25% throttle opening, and 35%, for wide-open throttle.
From the moment the refrigeration system was shut down the relative humidity showed a slight rise, and then, rose up at a high slope to values between 91% and 96%, for all valve openings. That is explained by water vapor being formed from ice melting with temperature increase after refrigeration system shut down, thus increasing air humidity.
7.3. Time variation of cooling capacity
Fig. 5. Time variation of cooling capacity
In Fig. 5 is shown the time variation of the absorption refrigeration system cooling capacity. Cooling capacity increases from the very beginning of the refrigeration system start up until reaching a maximum when the steady state condition is attained in the refrigerator interior, dropping from then on. Unlike the previous figures, here the results are shown until refrigeration system shut down.
The maximum cooling capacity obtained for the tested refrigerator was between 14.9 W, for 25% throttle valve opening, and 18.4 W, for wide-open throttle. However, the engine exhaust gas thermal power availability shown in Fig. 6, calculated by Eqs. (14)–(16) but considering the exhaust gas would leave the heat exchanger at the ambient temperature, presents similar values to that obtained by those authors. For determination of the coefficient of performance, heat transfer from the exhaust gas was calculated with the outlet enthalpy evaluated at the measured exhaust gas temperature leaving the heat exchanger (Eqs. (14)– (16)).
7.4. Engine exhaust gas power availability
Fig. 6. Engine exhaust gas power availability.
From Fig. 6 it can be inferred that the available exhaust gas power increases with wider valve opening. Power availability for 25% valve opening corresponds to approximately 18.8% that of 75% valve opening, and 18.4% that of wide-open valve power availability.
These values characterize an exponential growth of the exhaust gas power availability as a function of engine valve opening. For 25% valve opening, the system shows the lowest cooling capacity (Fig. 5) and the highest COP (Fig. 7). For wider valve openings the cooling capacity increases and the COP decreases, due to higher exhaust gas power availability. The present results showed that there is no need to use all the available exhaust gas energy to operate the absorption refrigeration system for throttle valve openings wider than 25%. In such situations, a control system for the exhaust gas mass flow rate may be a useful tool to optimize the absorption refrigeration system operation for automotive application.
7.5. Time variation of refrigerator coefficient of performance
Fig. 7. Time variation of refrigerator coefficient of performance
Fig. 7 shows the time variation of the refrigerator coefficient of performance until system shut down after the steady state condition was attained. The maximum coefficient of performance was 4.9%, for 25% open throttle. For all other valve opening, the peak coefficient of performance attained was even lower, of about 1.2–1.4%. This result indicates that a dedicated system is necessary to take full advantage of the energy available in the exhaust gas (see Fig. 6) and improve the coefficient of performance to reasonable values.
Figs. 8 and 12 show the influence of the absorption refrigeration system installed in the engine exhaust on performance and emissions parameters. Results are presented for engine output power, specific fuel consumption, and carbon dioxide, carbon monoxide and hydrocarbon emissions. These results should be observed considering that, to introduce the refrigeration system the engine exhaust system was modified, as mentioned in the previous section.
7.6. Influence of refrigerator on engine performance
Fig. 8. Influence of refrigerator on engine power output. Fig. 9. Influence of refrigerator on specific fuel consumption
Fig. 8 shows that, with the refrigerator installed in the exhaust system the engine produced 20% more power in comparison to the original configuration, corresponding to 2.2 kW at 25% throttle opening. On the other hand, the specific fuel consumption, that is, the fuel mass flow rate per unit power produced, was reduced for all valve openings, reaching up to 15% reduction at 25% valve opening (Fig. 9). These results show that the pressure drop introduced in the exhaust system by the presence of the refrigeration system is low, in comparison to the other exhaust system components, which were removed. As the exhaust pressure was lower, the power required from the engine to expel the exhaust gas was reduced, thus increasing the available output power. As the fuel mass flow rate into the engine did not change substantially, the higher power output decreased the specific fuel consumption with the refrigerator installed.
7.7. Influence of refrigerator on carbon oxides emissions
Fig. 10. Influence of refrigerator on carbon dioxide emissions Fig. 11. Influence of refrigerator on carbon monoxide emissions
Figs. 10 and 11 show that carbon dioxide emissions increased and carbon monoxide emissions decreased in the presence of the refrigerator for all engine operating conditions tested, except by 25% valve opening, for which the trends were the opposite. For 50%, 75% and wide-open throttle the mass fuel consumption was slightly lower in the presence of the refrigerator, being the reason for the reduction of CO emissions. For 25% valve opening a higher fuel mass flow rate was recorded when the refrigerator was installed, being the reason for the increase of CO emissions at this condition. Fuel mass flow rate can be obtained through multiplying the engine output power (Fig. 8) by the specific fuel consumption (Fig. 9).
7.8. Influence of refrigerator on hydrocarbon emissions
Fig. 12. Influence of refrigerator on hydrocarbon emissions.
Exhaust hydrocarbon emissions were consistently higher when the refrigerator was installed in the engine exhaust for all tested conditions (Fig. 12). This is a consequence of lower pressure attained in the engine combustion chamber during the exhaust process, as the exhaust system pressure was lower. The lower exhaust pressure increased the unburned fuel mass flow rate out of the combustion chamber crevices and from the cylinder lubricating oil film into the burned gas, thus increasing unburned hydrocarbon formation.
• The engine exhaust gas was confirmed as a potential power source for absorption refrigeration systems.
• The domestic absorption refrigerator tested showed low coefficient of performance and did not provide the cooling capacity needed for automotive application. However, a dedicated absorption refrigeration system may be able to take advantage of the exhaust gas power availability and provide the cooling capacity required for automotive air conditioning.
• Introduction of the absorption refrigeration system in the engine exhaust system did not cause significant pressure drop in the exhaust flow, as the engine output power was increased and specific fuel consumption was decreased with removal of other exhaust system components.
• Overall, carbon monoxide emission was decreased when the absorption refrigerator was installed in the exhaust gas, while hydrocarbon emissions showed an increase. Changes in exhaust components concentration were a consequence of the major modifications in the exhaust system.
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