Energy Library

Crude Oil Shale and Fuel gas

The present prime concern of politicians is no longer the rapid depletion of the finite reserves of non-renewable combustible resources, such as crude oil, natural gas and coal, but rather the adverse environmental consequences of their utilisation. Oil-shale is one of the largest relatively undeveloped natural, fossil-fuel resources in the world and so an important potential source of energy. It is, in some respects, similar to coal, being a highly-variable solid material with respect to its physical and chemical properties, which dictate the overall processing economics of a commercial-scale plant for its utilisation.

Production of synthetic fuels and/or electricity from oil-shale usually results in discharges of pollutants to the environment. (The waste from a processing plant is often not an unavoidable result of plant operation but a measure of its inefficiency: in general, the more efficient a plant is, the more unusable by-products it produces). The challenge facing the establishment of a commercially-viable oil-shale industry is not only the development of a simple, efficient and low-cost process, but limiting these adverse environmental impacts. This will require an understanding of various technologies for harnessing the energy from the oil-shale, such as mining; its preparation and crushing; retorting; direct combustion; as well as the disposal of spent shale and other wastes in an environmentally-wise manner.

Existing oil-shale deposits world-wide are abundant compared with other remaining fossil-fuel reserves. The availability of crude oil or natural gas can be measured in decades, whereas the identified readily-available oil-shale reserves are sufficient to satisfy the world’s energy needs for several centuries. This should ensure the further development of oil-shale (as an energy source at competitive cost) especially for electric-power generation.

Oil-shale conversion systems are presently at various stages of evolution. They range from those already being used in Estonia, Russia and China, to those currently being tested via pilot (or laboratory) scale projects in Japan, Australia, Morocco and the USA, to those that have been formulated conceptually (but still need to be tested), such as the proposed OSITGS. At present, oil-shale utilisation technologies are limited to either destructive distillation (i.e. retorting) processes to produce shale oil and synthetic gases, or direct combustion for electric-power generation and other industrial purposes.

There are major difficulties facing the development of the oil-shale industry, such as the environmental impacts of the processes involved. For example, in Estonia, employing pulverised oil-shale combustion systems incurs serious operational difficulties, including the low availability of the boilers as a result of corrosion, fouling and slagging (which are enhanced in the presence of alkali metals, sulphates and chlorides in the raw shale) as well as water-, land- and air-pollution problems. The average thermal-efficiency of existing commercial pulverised oil-shale plants is 30% (even without flue-gas clean-up) and they have low availabilities (i.e. 50%). In general, such technology for firing oil-shale is financially discouraging as well as environmentally unacceptable and hazardous. Thus, it has not achieved significant market-penetration world-wide. More stringent emission-control requirements and the need to generate electricity (or heat for industrial processes) from fossil fuels (at higher efficiencies and lower costs) have led to new technologies being introduced (e.g. fluidised bed-combustion, FBC). This is considered to be an energy-efficient and environmentally-friendly means of burning low-grade, high sulphur-content, low calorific-value solid fuels, such as tars; industrial, agricultural and municipal wastes; poor-quality coals; and oil-shale.

Retorting processes (based on heating oil-shale to relatively-high temperatures of 500°C) tend to yield many aromatic-hydrocarbons, which can combine to form carcinogenic compounds, such as polycyclic aromatic hydrocarbons. This significant health hazard, together with the necessity for the disposal of vast amounts of retorted shale resulting in groundwater contamination, can lead to serious adverse environmental impacts. It appears that most current technologies (except FBC) for utilising oil-shale still pose unanswered challenges.

In this preliminary study, a broad-brush analysis is undertaken deliberately because of the uncertainties concerning the future performances and costs of technologies that will be incorporated into the proposed system. More importantly, the main aim is to draw attention to the expected environmental impacts of the proposed OSITGS and appropriate mitigation measures that can be imagined at present.

The main objective of developing the OSITGS is to reduce Jordan’s dependence on imported crude-oil. This requires the development of an economically-competitive and environmentally-acceptable oil-shale industry, whose products can compete commercially without government subsidies. The proposed basic-plant is a complete facility for mining, retorting, gasification and combustion of the oil-shale, as well as for disposing of the spent shale. The preliminary design of this integrated system is based on proven technologies, such as oil-shale mining and retorting, gas-turbines, waste-heat recovery boilers, circulating fluidised-bed combustors (CFBCs) and ash cooling-and-disposal equipment.

The proposed plant will, most likely, be located close to the vast naturally-occurring oil-shale deposits. It would be capable of processing oil-shale into synthetic fuels and electricity, with a nominal output of 8000 barrels per day of shale oil and 400 MW_e of electric-power (as an installed generation capacity). It consists of the following.

Surface (i.e. strip or open-pit) mining will be used. The overburden and the oil-shale are loosened by explosives, and then extracted by draglines and/or shovels. The coarse oil-shale is crushed and transported away from the mine to the processing plant.

These receive and stockpile the crushed shale from the mine and provide surge storage between the mine and the processing units. The coarse shale is crushed and screened to the required size, then separated and stored in various hoppers (which feed the different processes within the plant) according to size.

The gasifier operates at relatively-high temperatures in order to produce the desired low-calorific-value (LCV) fuel gas, which is used to fuel the simple combined-cycle unit. Char from the bottom of the gasifier is fed, together with fine particles collected from the hot fuel-gas clean-up stream, to the CFBC.

The indirectly-heated retort produces shale oil and medium-calorific-value (MCV) fuel gas. Oil vapour and gas are collected and removed from the retort, in order to be cooled and separated. The retorted shale is circulated to the CFBC, so all the combustibles remaining are utilised.

Fine (of average size <6 mm) oil-shale particles and solid waste-streams (e.g. retorted shale and residual char), as well as gaseous and the difficult-to-dispose-of liquid by-products from other processing units, are fed to the CFBC. Superheated high-pressure steam is generated and used to drive a turbine generator to produce electric-power, and to supply the steam required for other processes. Fuel gas produced (from either the retort or gasifier) may be used as a supplemental fuel for ensuring the CFBC’s flame-stability.

These include the raw-water treatment unit, coolers, waste-water treatment plant, spent-ash cooling and disposal system as well as storage and general facilities.

The proposed plant should be established near the oil-shale deposits (i.e. in the central part of Jordan, near Karak), in the region extending from 100 to 150 km south of Amman. The local topography is flat to rolling with some hills: the average elevation is between 700 and 800 m above sea level. The climate there is very hot, dry and dusty in summer and cold and dry in winter, with monthly-average temperatures of 5°C in winter and 37°C in summer. However, the maximum temperature during the summer usually exceeds 40°C. The annual rainfall is normally between 50 and 100 mm, but the amount may vary significantly from year-to-year. Rain storms are localised and floods are comparatively few: hence, it is considered to be a semi-desert area. But there are a few small, shallow dams, which are filled during the rainy season but dry-up towards the end of summer. The prevalent wind is from the south west at 30 (±5) km/h during winter, but occasionally afternoon-wind speeds may reach up to 50 km/h. Whereas the summer wind is predominantly from the north west at lower speeds (i.e. 15 (±5) km/h).

Tags: oil shale

Posted in Applied Energy |

District Heating and Cooling Energy Equilibrium Model

Since many energy policies, e.g. strategies for utilizing new energy technologies, may have long-term economic impacts, many energy-related economic models have been developed to aid in energy planning and decision-making. In one class of energy-related economic models, the effects of energy policies are modelled as shifts in energy market equilibrium positions. Such models employ mathematical programming, and are often referred to as energy equilibrium models.

An energy equilibrium model of a competitive energy market examines the interaction between energy supplies and demands, and determines the optimal levels of production (supply) and consumption (demand) that satisfy the equilibrium property that the prices consumers pay for each commodity should equal the marginal costs of production. In the energy equilibrium model, supply is represented by a cost-minimizing linear submodel and demand by a smooth vector-valued function of prices. Several algorithms exist for the solution of the models, including the well-known Project Independence Evaluation System (PIES) algorithm of Ahn and Hogan.

In multi-period energy equilibrium, environmental impacts can be important factors, and an analyst may wish to introduce environmental measures into the model system. These environmental measures can be used to evaluate environmental impacts over the time horizon of the model for a given energy policy, or can be constrained by applying upper bounds to limit environmental impact. In the latter case, the model can provide the optimal solution of the model system while accounting for environmental impact control, e.g. CO₂ emission control.

Cogeneration-based district energy (DE) systems use central cogeneration plants with heating and/or cooling networks to provide electrical and heating and/or cooling services to communities. Such systems often have many advantages over conventional separate systems for electricity, heating and cooling, including increased efficiency, reduced environmental emissions and more economic, safe and reliable operation. Although the number of cogeneration-based DE systems is relatively small at present, the utilization of such systems is growing. One of the main impediments to their wider application is lack of experience with, and understanding of, the behaviour of integrated forms of such systems, which can often be complex and confusing. A larger base of knowledge exists for each of the component technologies when applied independently. The authors feel that the utilization of district energy and cogeneration could be greatly increased if better optimization and analysis tools were made available. The present work is intended to address this need. More specifically, we employ in this paper an energy equilibrium model to study conventional heating, cooling and electricity-generation systems and cogeneration-based DE systems in order to compare the potential economic and environmental benefits of utilizing cogeneration-based DE systems for several scenarios, and to develop the optimal configuration of the model system, considering such factors as economic and environmental impacts.

The energy equilibrium model is set up, formulated and solved within the software called the Waterloo Energy Modelling System (WATEMS), which employs sequential nonlinear programming to calculate a spatial intertemporal equilibrium of energy supply and demand.

Section 2 of this paper discusses cogeneration and district energy, while Section 3 presents the energy equilibrium model and its mathematical formulation. Section 4 explores methodologies for analysis and evaluation. In Section 5, an illustrative case study is presented. Section 6 presents the conclusions.

A cogeneration process involves the simultaneous production of electricity and heat (usually in the form of steam and/or hot water). The main advantage of cogeneration is that less input energy is consumed than would be required to produce the same thermal and electrical products in separate processes. Additional benefits of cogeneration often are reduced environmental emissions (due to reduced energy consumption and the use of modern technologies in large, central installations), and more economic and safe operation. The additional safety associated with cogeneration in part is due to the fact that only one fuel-fired combustion plant is required, compared to the two such plants needed for separate heating and electricity-generation systems. Cogeneration has been used, particularly by industry, for approximately a century, and there are presently over 4000 cogeneration projects. Cogenerated heat can satisfy air- and water-heating demands in the residential, commercial as well as institutional sectors (using on-site cogeneration, or central cogeneration with district heating, and industrial heating needs (e.g. drying, boiling). Cogenerated heat can also provide space cooling via heat-driven absorption chillers.

Cogeneration systems are similar to thermal electricity-generation systems. In most thermal electricity-generation systems, an energy resource (normally a fossil or nuclear fuel but sometimes a renewable energy resource) is converted to heat, of which a portion (normally 20 to 45%) is converted to electricity, and the remainder rejected to the environment as waste heat. In cogeneration systems, depending upon the needs of the customers, part of the generated heat is used for electricity production, part is delivered as product, and waste-heat output is reduced. Cogeneration energy efficiencies (based on both electricity and heat) of over 80% are achievable.

District energy systems (which can include both district heating and district cooling systems) use central heating and/or cooling facilities to provide heating and/or cooling services for communities. In a district-cooling system, a chilled fluid, normally treated water, is supplied from a central chiller plant and transported by pipeline to users of the cooling capacity, then returned for recooling. The chilling plant can utilize electrical chillers or heat-driven absorption chillers. In district heating systems, a similar heating loop with a central heat supply is utilized. The advantages of district energy systems over conventional heating and cooling systems include improved efficiency, reliability and safety, reduced environmental impact, and for many situations better economics. The increased safety associated with DE systems exists because the heating and cooling plants, where problems are likely to originate in the event of accidents, are located at a different site than the user of the heating and cooling, where DE is used. The increased reliability of DE systems is attributable to the fact that many DE users are normally connected through a network having more than one heating and cooling provider; thus in the event of a breakdown at one heating/cooling location, an alternate can usually be started up to avoid service disruptions. District energy systems can be particularly beneficial when integrated with cogeneration systems.

Many integrated systems for cogeneration and district energy are possible. Several selected applications are discussed below:

• A district-heating and cooling system operated by Energy Networks Incorporated since 1962 currently serves over 70% of the buildings in downtown Hartford, Connecticut. A natural gas-fired cogeneration plant, completed in 1990, produces the hot and cold water required for the system, as well as electricity.

• The feasibility of district heating and cooling was assessed for downtown Des Moines, IA, considering a mothballed 210 MW (electric) coal-fired power plant as the source of heat. The system was predicted to break even economically in 20 years, and have a lifetime of 40 years. District cooling, using electrically-driven centrifugal chillers, was estimated to provide cooling at competitive prices. The use of absorption chillers driven by hot water from the district heating system was estimated to be slightly more expensive.

• Edmonton Power proposed for downtown Edmonton, AB a major cogeneration-based district heating and cooling project having (i) an initial supply capacity of 230 MW (thermal) for heating and 100 MW (thermal) for cooling, with the potential to expand to about 400 MW (thermal) for heating over the next 10 years; (ii) the capacity to displace about 15 MW of electric power used for electric chillers through district cooling; and (iii) the potential to increase the efficiency of the Rossdale power plant, which would cogenerate to provide the steam for district heating and cooling, from about 30 to 70%. Electrical chillers were to be used originally, and absorption chillers in the future.

Tags: heating and cooling

Posted in Applied Energy |

Recovery and Occupancy Period Point Temperature

The economic interest of intermittent heating systems in intermittently-occupied buildings (e.g. schools, office buildings) is no longer in doubt. Nevertheless, optimal energy consumption requires that the heating restart time be defined with great precision. In many cases, for fear of ill evaluation, the lowering of the temperature during non-occupancy period is minimal, which leads to reductions in energy savings.

The aim of this article is to compare the precisions of the different methods for calculating heating restart times, and to study their impacts on comfort and on energy consumption. Several classical methods are studied. For example, the method where the system is restarted at a fixed time and the methods for which the recovery duration is analytically calculated according to internal and/or external temperatures. We also propose two new methods. The first of these determines the duration of the recovery period according to the external and internal temperatures, by means of fuzzy logic. The second is based on the use of a two time-constant building model. Apart from the presentation of these two new methods, the originality of the study is that it carries out the comparison using a heating law, during occupancy periods, adapted to intermittent heating.

An intermittent heating controller allows the internal temperature to be lowered during non-occupancy periods, while maintaining the desired temperature during occupancy periods. The different phases of heating are as follows:-

• upper control during occupancy period. The internal air temperature Ta must be maintained at the upper set-point temperature Tu by means of the heating law.

• minimum power: switching-off of heating at the end of the occupancy period.

• lower control during the non-occupancy period if the internal temperature reaches the lower set-point temperature Tl. The lower set-point temperature avoids the risk of condensation or frost.

• recovery at maximum power, so as to reach the upper set-point temperature Tu from the beginning of the following occupancy period. Restart time is optimized by the controller.

As far as restart optimization is concerned, the controller estimates at each time step, throughout the non-occupancy period, the recovery duration, Dur, which corresponds to the duration needed to reach the upper set-point temperature. When this period is equal to the period remaining until the next occupancy-period, the heating is restarted automatically. If the recovery period continues into the occupancy period, the comfort of the occupants is not ensured. On the contrary, if the start-up is anticipated, there will be a corresponding loss of energy. It is therefore necessary that duration, Dur, calculated before the restart, should correspond as closely as possible to the real recovery duration, Durr, which is observed a posteriori.

The simplest solution consists in determining a fixed duration, by distinguishing Monday from the other days of the week. This solution may be adequate as far as regulations are concerned, but it is imprecise because no account is taken of the climatic conditions.

Seem et al. has shown that the duration of the recovery period depends essentially on the internal temperature at the beginning of the recovery period. But, when recovery time is long, climatic factors are more important and it is thus necessary to take the external temperature into account.

The use of a building model within the controller allows us to evaluate the duration of the recovery period. The precision is of course dependent on that of the model. Different types of models are possible. For example, Visier et al. consider a physical model for optimal control. A neural network is used by Glorennec to model the coupling of a building and an underfloor-heating system. Finally Botte considers an Armax type model.

In this article, we will use a two time-constant building model. This model has the double advantage of being both simple and sufficiently accurate to correctly estimate recovery duration.

Fuzzy logic is based on the theory of fuzzy subsets, first introduced in 1965 by Professor Zadeh of the University of California. In this theory, an element belonging to a set (a fuzzy subset) is between 0 and 1. It is not equal to either 1 or 0 as is the case in classical logic.

Fig. 2 presents a breakdown of the Te discourse universe into three fuzzy subsets: Low, Medium and High (two trapezes and one triangle). When variable Te is equal to 2°C, the degree of membership to each of these fuzzy substets is respectively 0.8, 0.2 and 0.

The controller developed in this study allows us to define different optimization laws in order to compare them.

The classical law determines the supply power in relation to the external temperature so as to compensate for the building’s steady-state thermal losses. Thus it is not entirely suitable for intermittent heating since the energy consumption needs are greater on Monday morning after a weekend set-back, than for other days of the week. This is why we use a surface heating law, based at once on Te, and on a State variable, which takes into account the thermal state of the building.

We considered six cases corresponding to different methods of optimization. The duration of the recovery period Dur can be calculated either directly from internal and/or external temperatures (cases 1 to 5), or indirectly, repeatedly by means of a building model (case 6). The various cases studied are based on classical analytical functions, with the exception of case 5 which is based on fuzzy logic. Cases 5 and 6 constitute new methods of determining heating-restart times.

A learning mechanism based on the recursive least squares routine automatically modifies the three parameters defining each of the two laws. The use of a recursive method, as opposed to a direct method, avoids stocking too great a number of parameters in the controller.

Tags: occupancy period

Posted in Applied Energy |

Emission Factor of Sulphur Dioxide Content

Part 1 in this series dealt with the emissions of carbon dioxide: many of the factors that influence the emissions of carbon dioxide also apply to sulphur dioxide. As with carbon dioxide in Part 1, the ‘emission factors’ for sulphur dioxide have been derived by direct calculation involving fuel usage and composition.

The AEA Atmospheric Emission Inventory Annual Report provided a major source of data for this paper. It also provided an ‘estimated precision value’ associated with calculating sulphur-dioxide mass emissions: it is of the order of ±10% and is largely dependent on the accuracy of the monitoring of the consumption, fuel analysis and estimations of the sulphur retained in the ash.

The ‘emission factors’ established for the various types of fuels, coupled with the historical trends of changing patterns of energy utilisation in the UK, have been used to establish the corresponding changing pattern of sulphur-dioxide mass emissions over the 1970–1994 period. The changing demands for primary energy have been discussed in Part 1 of this series.

These data show the overwhelming dominance of the energy sector as the major source of sulphur dioxide and clearly demonstrate that, over this period, 99% of sulphur dioxide is generated by the energy sector. It also demonstrates that solid fuels were responsible for 75% of these emissions and oil products the remainder.

The DOE report has shown that the annual mass emission of sulphur dioxide, arising from fossil fuels, decreased from 6.37 Mtonnes in 1970 to 2.7 Mtonnes in 1994, and this paper aims at providing a detailed analysis of this decline.

The calculated theoretical SO₂ emissions in grammes of SO₂/kg for each of the fuel types outlined in the following table are based on the fuel consumption, heat content and on an estimate of the retained sulphur in the coal ash estimated to be 20% in the domestic sector and 10% in the industrial and commercial sectors.

Natural gas consumed in the UK contains a very low sulphur content and, in this analysis, is regarded as having an emission factor of zero.

It can be seen from Table 2 and Table 3 that the emission factors for coal are spread over the wide band from 640 to 1424 grammes/GJ corresponding to 1–1.7% sulphur content, whereas the range of emission factors for fuel oil is much narrower, i.e. 1149–1305 grammes/GJ with sulphur contents of 2.5?2.8%. The distillates are at the lower end of the emission factor spectrum of 18–178 grammes/GJ with sulphur contents ranging from <0.05 to 0.4%.

Solid fuel was the major contributor to the mass emission of sulphur dioxide in 1970 and retained its number-one position in 1994, while natural gas was a zero contributor to sulphur dioxide in 1970 and 1994.

Clearly there has been a major overall downward trend in the level of SO₂ emissions from the fossil fuels. The replacement of coal and petroleum by natural gas over this period has been the major reason for the decline in emissions by all the users with the exception of transport, where there has been no fuel substitution and the demand for energy in this sector has expanded rapidly.

The corresponding percentage contribution of sulphur dioxide generated compared with the changes in fossil-fuel consumption for the years 1970 and 1994 are shown in Table 5.

The downward trend in the demand for solid fuel is reflected in the overall decline in the mass emissions of SO₂ arising from this fuel.

The decline in sulphur-dioxide emissions arising from petroleum are attributed to the reduced demand for this product coupled with the reduced consumption of high sulphur-bearing fuel-oil by the manufacturing and power-generating industries and the dramatic increase in the demand for the light low-sulphur bearing fuels used in the transport sector.

Globally, SO₂ emissions from fossil-fuel combustion have been seen to increase in the 20th century: this trend has been most evident in industrially developing countries during the 1970 to 1994 period. The annual SO₂ emissions in the UK, as shown in Fig. 1, declined from 6.4 to 2.7 million tonnes, i.e. a 58% reduction, which is equivalent to a 2.4% annual fall. The emissions peaked in the early 1970s, but have shown a significant downward trend ever since. It can be seen from Fig. 1 that the emissions arising from solid fuels remained relatively stable up to 1987, but declined thereafter. However, the emissions from petroleum steadily declined from 1970 to 1984 and were stable thereafter, which can be explained in both cases by the change in the energy-mix employed largely in favour of natural gas. Fig. 2 shows the decline in the SO₂ emission/energy unit (emission factor) indexed to 1990=100: it amounted to a 53% fall. The relationship between annual SO₂ emissions to GDP calculated relative to the value at 1990 is demonstrated in Fig. 2 in that, during 1970 the SO₂ emission was 18.4 tonnes per £ million of GDP at 1990 price, but by 1994 this had declined very dramatically to 5.5 tonnes per £ million of GDP. This could be accounted for by the reduced energy intensity in the manufacturing industry and the change in fuel mix to low-polluting fuels.

Fig. 3 demonstrates the changing pattern of the estimated SO₂ emission-factors from 1975 to 1993. The factor for coal remains relatively stable over this period; however, there is a marked decline in the petroleum emission-factor over the 1975 to 1990 period. This is largely due to the dramatic reduction in the demand for the high-sulphur fuel-oil consumed by manufacturing industry and the increase in the demand for road-vehicle fuel, which has a low sulphur-content compared with industrial fuel-oils.

Fig. 4 outlines the estimated mass emission of sulphur dioxide arising from the individual energy-consuming sectors based on the product of the calculated or empirically-derived emission factors and annual fossil-fuel consumptions. It identifies the changing pattern in the emissions that have occurred over the period 1970 to 1994 with the major energy-consuming sectors. It can also be seen that in the power, domestic/service and industrial sectors, sulphur-dioxide emissions have declined.

Four major pieces of environmental legislation have been enacted to control sulphur dioxide during the 1970–1994 period.

Firstly, the Control of Pollution Act of 1974 included regulations under section 76 to limit the sulphur content of oil fuel (defined as any liquid-petroleum product produced in a refinery); the later EC Directive 87/219 EEC set a limit of 0.3% on the sulphur content of gas oil used as a domestic fuel.

Secondly, in 1980, a European Council Directive 80/779/EEC was enacted and came into force in 1983: it prescribed mandatory maximum values for ground-level concentrations in ?g/m³ of smoke and sulphur dioxide.

Thirdly, an EEC Directive has been enacted imposing legislative limits on the sulphur content of petrol and diesel fuels. In 1990, the standards were 0.1% for petrol and 0.3% for diesel fuel. In 1994, the legislation reduced the permissible sulphur content for diesel fuel from 0.3% to 0.2% and by 1996 both petrol and diesel fuel had a maximum of 0.05% sulphur. The road-transport sector only contributes 4% of the total UK sulphur: however such emissions are generated at ground level and not dispersed via tall chimneys and so is very relevant to public health.

Fourthly, in 1990, a European Council Directive was enacted to limit the sulphur dioxide emissions into the air from large combustion plant with thermal inputs equal to or greater than 50 MW. In response to this Directive, the UK prepared a national action plan for the reduction of SO₂ emissions from these plants for the period 1991?2003. The plan was designed to limit the overall emissions from these plants, from a base line in 1980, by 15% in 1987, 21% in 1993, 45% in 1998 and 63% by 2003.

Tags: sulphur dioxide

Posted in Applied Energy |

Weak Solution and Stage Absoption Heat

To increase the performance of absorption heat transformers, there has been a continuous effort to obtain high temperature lift and COP by using the exhaust gas as the heat source. A two-stage absorption heat transformer (TAHT) is one kind of advanced heat transformer in which it is possible to achieve a higher absorber temperature. A recent study has presented a mathematical model of a two-stage heat transformer with a solution heat exchanger operating with water/sulphuric acid. Rivera et al. has proved that the double configuration can have a higher absorber-temperature, but the COP is less than that for the single-stage absorption heat transformer. Grossman developed a simulation program called “A user-oriented computer code (ABSIM)” for a multi-stage absorption system, which can be employed to investigate various cycle configurations with different working-fluids.

An exergy analysis and the concept of the second law of thermodynamics play an important role in evaluating the behaviors of thermal and chemical systems, because their applications lead to a better understanding of energy transformation processes. Energy utilization diagrams (EUDs) represent features of both the first and second laws of thermodynamics with respect to energy transformations. The pinches can readily be observed on these EUDs.

In the present paper, a modification of a two-stage absorption heat transformer with the incorporation of the latent and the sensible heat modes is implemented. We discuss the case in which the exhaust gas from a practically-operating gas-turbine plant is taken as the heat source for the TAHT. The results show that this TAHT is more efficient, the exergy efficiency is increased from 48.14 to 54.95% and the temperature lift is increased by 10.7 K.

Two-stage absorption heat transformers basically consist of two generators G1 and G2, two condensers C1 and C2, an evaporator E1, an absorber evaporator A1 and an absorber A2, as shown in Fig. 1. Heat from a source is divided into two streams. One of them is supplied to separate water from a water/lithium bromide mixture in generator G1. The vaporized water is condensed in the condenser C1 and then is pumped into evaporator E1, where it is vaporized at an intermediate temperature and pressure P_H1. The other stream enters the generator G2. The generated water vapor is condensed in condenser C2; then the condensed water is pumped, at pressure P_H2 which is slightly higher than P_H1, and vaporized by receiving heat Q_A1. The vaporized water is absorbed in absorber A2 at a high temperature by the strong solution coming from generator G2. The weak solution, via the heat exchanger HEX2, enters generator G2, so repeating the cycle again.

Fig. 2(b) displays the arrangement of the generator in close-to-equilibrium operation. We keep the pressure nearly uniform and divide the generator into, say, four compartments. Because the temperature of the weak solution that comes from the solution heat exchanger is high, a large temperature-difference would exist. If this stream enters the first compartment (i) directly, a large exergy-loss will be generated. Hence, the weak solution stream releases heat during its passage through compartments (iv, iii, and ii) and finally flows into compartment (i). The superheated vapor evaporated from each compartment releases heat in the upper compartments, joins with the vapor stream from compartment (i), and then enters the condenser. Also the external heat-source gives evaporation heat to the weak solution in compartments (iv) to (i). In compartment (iv), the concentration of the weak solution reaches the required concentration and is fed through the heat exchanger into the absorber. By selecting proper temperatures for compartments (i) to (iv), close-to-equilibrium operation can be achieved. For comparison, we include a single-compartment generator in Fig. 2(a).

Fig. 3(b) shows the configuration of the absorber. We split the process into four absorption-processes. Namely, the saturated vapor is divided into four streams, one to be introduced to each compartment. Meanwhile, the strong solution is heated in compartments (iii) and (ii) and enters compartment (i). The heat released by absorption is accepted also by the feed water in compartments (i) to (iv). The strong solution passes compartments (i) to (iv), becoming dilute, and enters the solution heat exchanger. To compare this scheme with a single-compartment absorber, we introduce Fig. 3(a).

As can be seen in Fig. 4, we can compare the one-compartment absorber with the multiple-compartment absorber by introducing the Dühring diagram. To express clearly and simply, we explain this change in a single-stage absorption heat transformer (SAHT). Fig. 4(a) shows the Dühring diagram of the single compartment absorber. Because the temperature of the pure saturated vapor is low, the final point in equilibrium with the weak solution is not at point H, but is A, as shown in Fig. 4(a). As a result, point H is an imaginary state that cannot be attained. If we make the state point A close to point H, we can obtain a higher temperature-lift.

Based on this consideration, the multiple-compartment absorber in Fig. 4(b) is proposed on the basis of the sensible-heat mode. Its Dühring diagram is shown in Fig. 4(b). Here, the strong solution enters the absorber gradually by absorbing the split-saturated vapor at two different pressures. Adopting two different pressures, P_H1 and P_H2 in Fig. 1 can make the temperature of useful heat increase, namely, A1 approaches the imaginary point H much more closely.

Consequently, the strong solution passes through equilibrium points A1, A2, A3 and finally gets to point A in equilibrium with the weak solution. The highest temperature of the absorber is seen at state A1. A similar phenomenon also occurs in the multiple-compartment generator. We apply this multiple-compartment concept to absorber A2 and generators G1 and G2 in Fig. 1 (case II).

Fig. 5 shows the improved configuration of a TAHT. The water vapor, which comes from evaporator E1, is split into two streams, one flows into compartment (i) and the other enters compartment (ii) in absorber A1. The strong solution that comes from generator G1 flows into compartment (i), where it absorbs the vapor from evaporator E1, becomes dilute and then enters compartment (ii) of absorber A1. In compartment (ii), this solution absorbs the vapor from evaporator E1, becoming the weak solution, and enters the solution heat exchanger. The liquid water from condenser C2 is divided into two streams and is pumped to different pressures. The first stream is heated by absorption heat in compartment (i) and the other stream is heated in compartment (ii). Finally, these two saturated vapor streams are introduced into the four-compartment absorber, i.e. absorber A2.

We use the calculation assumptions in Table 1. The thermophysical properties of aqueous lithium bromide solution are based on the ASHRAE Fundamentals data. To evaluate the performance of the absorption heat transformer system, two criteria, namely the coefficient of performance (COP) and the exergy efficiency are applied.

The performances of basic TAHT (case I) and the two improved versions (cases II and III) are summarized in Table 2. It can be observed that the total exergy loss is 299.1 kJ/s and the exergy efficiency is given by 48.14% for the basic THAT. By introducing the improved versions, the exergy loss will be decreased to 251.1 kJ/s in case II and 239.2 kJ/s in case III. Simultaniously, the exergy efficiency is increased to 50.99% in case II and 54.95% in case III. In addition, the heat-source temperature can be extended to 331.0 K. However, the COP is slightly reduced, say, to 30.83% in case II and 29.34% in case III but 31.44% in case I.

Tags: weak solution

Posted in Applied Energy |

Dichroic Disk in Infrared Visible Light

Natural ventilation and daylighting are increasingly employed in modern buildings, underground spaces and tunnels to minimise energy consumption and release of harmful emissions to the environment. Innovative daylighting techniques including lightshelves, prismatic glazing, holographic films and light pipes have facilitated the effective use of daylighting in a wide range of buildings. In addition to bringing energy savings, these daylighting technologies also help to create healthier interiors for occupants. Natural daylight has also been found to relieve seasonal affective disorder, cholesterol problems, chronic fatigue, jet lag as well as benefiting people in shift work and computer VDU work.Natural ventilation techniques have also been widely used and include passive stack systems in a wide variety of buildings across Europe. Until now, daylighting and natural ventilation techniques have been developed independently and form separate systems. Integration of these technologies would reduce system costs and payback periods as well as make natural ventilation and daylighting more attractive to owners and users of buildings.

One method for the integration is the use of concentric channels for both daylighting and natural ventilation. As shown in Fig. 1, the central channel, or the light pipe, will guide sunlight and daylight into occupied spaces while the outer channel, or the ventilation stack, enables passive stack ventilation. By constructing the light pipe using dichroic materials, the infrared part of the solar radiation is allowed to be transmitted to the stack but the visible light is reflected downwards within the light pipe towards the room interior. The heat gain to the interior can be reduced and the thermal stack effect strengthened.

A dichroic material is usually manufactured on a glass or plastic base in which alternate layers of transparent materials are laid. The amounts and values of the wavelength transmitted or reflected depend on the thickness and refractive index of each layer. Examples of dichroic materials include magnesium fluoride/zinc sulphide and tantala/silica oxides. The dichroic material used in the tests reflects the visible light while transmitting the infra-red radiation, at very high efficiencies. Thus, it is referred to as a “Cold Mirror”.

Experiments were set up to determine the infra-red and visible-light transmission/reflection characteristics of the material. This information then formed the basis of a detailed analytical study of the effect of dichroic material on light transmission in the light pipe and enhancement of air flow in the passive ventilation stack.

Tests were carried out in a specially constructed room with black internal surfaces to obtain the characteristics of the dichroic material in terms of infra-red and visible light transmittances and specular reflectances for visible light, as detailed below.

The dichroic material for testing was obtained from Sycamore Glass Components, USA. The material has a nominal reflectance of 90% between the wavelengths of 420 and 630 nm and a nominal transmittance of 85% on average between the wavelengths of 750 and 1200 nm. The 750–1200 nm band carries most infra-red energy in the solar radiation. The substrate is Borofloat, a borosilicate glass that can handle heat. The high cost of the dichroic material currently purchased from Sycamore precludes the construction and testing of actual light pipes lined with the material. However, recent advances in material and manufacturing technologies indicated that dichroic materials’ cost would soon be dramatically reduced to allow large-scale applications in buildings.

The light source used for the tests was manufactured by General Electric (GE) lighting. The lamp is Halogen TAL 100 mm with an integrated metal reflector and constant colour coatings providing consistent light quality and high intensity. The power of the lamp was 50 W and the beam had a peak intensity of 55,000 cd and beam spread of 4°. The colour temperature of the lamp is 3000 K and with a rated average life of 3500 h. The environmental chamber within which the optical experiments were carried out is a wooden structure measuring 3×3×3 m, with a single door and no other openings. The walls, door, floor and ceiling were covered in a black matt paper to reduce the amount of secondary reflection, which may affect readings, but only to a negligible amount. The test rig was on a platform inside the chamber and the equipment was operated from outside the chamber without the presence of the researcher inside the chamber. This also reduced the amount of unwanted secondary reflection. All experiments on the dichroic material were carried out under these strict conditions to ensure that accurate results were obtained.

The tests for the dichroic study were performed on a platform situated at the centre of the environmental chamber. Fig. 2 shows the schematic layout of the typical arrangement for the measurement of transmitted infra-red and visible light through a dichroic-coated glass disk. Steel rods and clamps were used to fix a 0.5 m long black tube horizontally on a table covered with matt black paper. The position and the horizontal levelling of the tube were checked regularly with a spirit level. The light source was placed at one end of the narrow black tube and the dichroic disk positioned on a spectrometer at the other end of the tube. The spectrometer allowed accurate adjustment and measurement of the angle between the dichroic disk and the light beam. The irradiance of infra-red and illuminance of visible light were measured both before and after the dichroic disk is positioned in place. These data were then used to calculate transmittance values.

The tube was coated on the inside and the outside with a matt black paint to ensure that the light rays meeting the dichroic disk are very close to being parallel. This allowed the accurate determination and adjustment of the incidence angle used in the tests. It was anticipated that, as the experiment continues, the tube would get hot due to the radiated heat from the lamp which may cause errors in the infra-red measurements. The lamp operating time and the corresponding extent of this effect were monitored and the information was then used to plan the test procedures. Measurements were taken during a short period of a few seconds between cooling intervals of about 15 min to control the temperature of the tube and to prevent the tube from heating up. This eliminated the error due to the radiated energy from the tube. The instrumentation was controlled and measurements undertaken from outside the chamber. The photometer used was a Hanger Universal Photometer/Radiometer model S3 fitted with sensors for the measurements of visible light illuminance and a special detector for the measurement of infra-red irradiance. It incorporated silicon-diode photocells with approximately the spectral sensitivity of the human eye (CIE). It was capable of measuring illuminance in the range 0.01–200,000 lux. It has an accuracy of ±3% and a virtually perfect cosine correction curve. The photometer was connected to a remote photocell via a flexible lead, making it easier to obtain readings without blocking incident radiation reaching the photocell. The infra-red remote sensor (SD7) had a spectral response to wavelength in the range of 700–1150 nm. It had an absolute sensitivity of 2 nA/W/m², with an accuracy of ±3%.

Tests to determine the visible-light specular reflectance of the dichroic disk were carried out in overcast-sky conditions. Prior to each test, a luminance range check was carried out to ensure that the overcast-sky condition was present at the time of the test. The overcast condition is assumed to occur when the average horizon luminance does not exceed half the luminance of the zenith. The tests were carried out in a large room, with large windows open to the outside, using the technique of Fontoynont and Berrutto. A piece of matt white paper (0.2×0.3 m) is attached to the window glass. The dichroic disk is positioned on the floor at a distance from the sheet of paper to allow light from the paper to strike the disk surface at a required angle of incidence on the disk to be established. The specular reflectance at a specific incidence-angle was determined by measuring the luminance “Lr” of the image of the paper reflected off the dichroic disk and the luminance from the paper, “Ls”. The ratio (Lr/Ls) of the two values is the specular reflectance of the dichroic disk at the specific angle-of-incidence. The test was repeated for a range of incidence angles of 10 to 80° at 10° intervals.

Tags: visible light

Posted in Applied Energy |

Screen Surface Air Vortex Diameter

Mechanical energy is produced when heat is carried upwards by convection in the atmosphere. The Manzanares solar chimney, shown in Fig. 1, was built in Spain in the 1980s and consisted of a vertical tube 200 m high and 10 m in diameter with a turbine installed inside its base. The chimney was surrounded by a solar collector, a transparent plastic roof 240 m in diameter supported 2 m above the ground. The air flowed through the open rim of the collector and up the chimney. The solar collector increased the air temperature by some 20°C; the upward velocity in the chimney was typically 10 m s^?1. The total insolation on top of the collector was 45 MW and the turbine generated 48 kW of electrical power for an overall efficiency of 0.1%. The operating conditions shown in Fig. 1 are based on those described in Ref. 2.

A power station using a controlled tornado-like vortex instead of a physical tube was proposed in Ref. 3. The centrifugal force in the annular vortex replaces the physical tube. There is also no need for a covered solar collector because the boundary layer acts as the solar collector. The vortex could extend from the earth’s surface up to the tropopause and the heat to work conversion efficiency could be 15%.

The proposed vortex power station is shown in Fig. 2. The vortex would be started by heating the air in a circular station with fuel while giving the air converging towards the centre of the station angular velocity by having the air pass through a rotating perforated screen. Once the vortex is established it would persist without fuel and its base would remain at the centre of the station. An optional vertical-axis turbine located in the centre of the station would generate electricity. The earlier proposal used fixed deflectors instead of a rotating screen. However, a rotating screen would provide more positive control during development.

A medium size vortex power-station could be 300 m in diameter, and the perforated rotating screen could be 50 m high. The vortex could be 80 m in diameter at its base. The turbine could produce 100 MW of electrical power from a vortex the size of a small tornado. The heat needed to sustain the vortex, after heating with the fuel has stopped, would be derived from the sensible and latent heat content of the air at the bottom of the atmosphere.

The process could be used to produce convective vortices ranging in size from dust devils to medium-size tornadoes. Tornadoes are dangerous, but the process could be developed safely by using physical models of increasing size, first indoors and then outdoors. Some dust devils are less than a metre in diameter. Under optimal conditions, it should be possible to start a self-sustaining vortex to demonstrate the concept with a station 30 m in diameter. Large stations could be tested safely by supplying heat continuously under stable atmospheric conditions.

Fire whirls have been produced by burning fuel in the centre of a rotating screen. The author has built a 30 cm diameter model—essentially a small-scale version of Fig. 2. The model consisted of a circular plate, with a 20 cm high vertical perforated-screen attached to its edge. The screen was an ordinary metallic bug-screen. The plate was placed on a turntable; there is no need to rotate the base, but on such a small model, it is simply easier to rotate both the screen and the base. The vortex was produced by burning liquid fuel on the base of the model while the screen was rotating. The fuel could be placed either in a circular cavity at the centre of the model or in an annular groove just inside the screen. The vortex stayed in the centre of the model. The vortex, which was 1–5 cm in diameter was stable and visible to a height of 1 m because of the colour of the flame and smoke. Some vortices extended to a height of up to 2 m. The model was only used indoors because small vortices are easily disturbed by stray air currents.

The air converging towards the base of the vortex is entrained sideways as it passes through the small openings of the screen and acquires a tangential component of velocity approaching the tangential velocity of the screen. The tangential velocity of the air just inside the screen was measured to be 87% of the screen tangential velocity by [4]. As the air converges from the rotating screen, its tangential component of velocity increases to conserve the angular momentum acquired at the screen, except in the layer adjacent to the surface, where tangential velocity is reduced by friction. As a result convergence is limited to the thin surface layer.

A very small turbine was installed in the centre of the model to test a method of extracting energy. The turbine was 4 cm in diameter and sat on the tip of a pin. For the turbine test, the fuel was burned in the annular groove located just inside the screen. The speed of rotation of the turbine was estimated at over 1000 rpm, i.e. much higher than the speed of rotation of the turntable, which ranged from 30 to 80 rpm. The flame would cling to the surface of the plate and impinge directly on the turbine. The turbine behaved like a cup anemometer caught in a rotating flow.

Fig. 3, where q is the heat received during processes 1–2, h is the enthalpy of the raised air including the enthalpy of its water content; g is the acceleration of gravity; z is the height of the tube, and v is the velocity. For an adiabatic process, (i.e. q=0), with negligible inlet and outlet velocities (v?0), the total energy equation reduces to

(2)W_b=??h??gz

The work is equal to the decrease in enthalpy of the air minus the increase in potential energy of the air. The work is a maximum when the process is frictionless and reversible, when the expansion is isentropic. The maximum work is, therefore, equal to the reduction in enthalpy minus the increase in potential energy in an constant entropy process, (i.e. s=constant).

The work due to buoyancy (w_b) is equivalent to the convective available potential energy (CAPE), which is widely used in meteorology, and which is the integral part of the force of buoyancy times the distance moved. During periods of insolation, CAPE is typically 1200–2200 J kg^?1. The average CAPE during a recent month of observation in an oceanic tropical area was 1920 J kg^?1. The maximum work of buoyancy is readily calculated from atmospheric soundings. The following oceanic tropical conditions will be used to demonstrate the technique by calculating the work produced when air is raised from the surface to the 20 kPa level. The conditions are: P₁=101 kPa; T₁=27°C; U₁=80% corresponding to m₁=18.18 g kg^?1; P₂=20 kPa and z₂=12400. In tropical oceanic areas, the level of neutral buoyancy is usually above the 20 kPa level and the elevation of the 20 kPa level is typically 12400 m. The corresponding energy variables are: h₁=73485 J kg^?1, h₂=?51954 J kg^?1, s₁=s₂=256.7 J kg^?1, ?h=125440 J kg^?1 ?gz=123730 J kg^?1. The work of expansion when the air is expanded from 101 to 20 kPa is 125440 J kg^?1, but 123730 J kg^?1 is required to lift a kilogramme of air including its water content to the 20 kPa level. The net work is therefore: w_b=1710 J kg^?1.

The effect of sounding properties is difficult to see from Eq. (2), but is readily appreciated by applying Eq. (2) to different conditions. Increasing the temperature of the surface air by 1°C at a constant mixing ratio increases w_b by 250 J kg^?1. Increasing the relative humidity of the surface air by 5% at a constant temperature increase w_b by 585 J kg^?1. Increasing the mixing ratio of the surface air by 1 g kg^?1 at a constant temperature increase w_b by 517 J kg^?1. Increasing the temperature of the surface air by 1°C at a constant relative-humidity increases w_b by 825 J kg^?1 because both the temperature and the mixing-ratio increase. A small change in the air temperature has a large effect on the work of buoyancy. Decreasing the temperature of the surface air by 2°C at constant relative-humidity, would reduce w_b to near zero. Decreasing the average temperature of the sounding by approximately 2°C, without changing the surface air conditions, decreases the level of the 20 kPa surface by 100 m and increases w_b by 1000 J kg^?1.

Tags: screen surface

Posted in Applied Energy |

Total Trip Length on Fuel Consumption

It is strongly required to enhance energy conservation to reduce fossil-fuel consumption to cope with global warming. In Japan, energy consumption in the industrial sector is almost stable after the oil crises. However, the commercial, residential and transportation sectors have a trend of increasing energy consumption. Especially reducing energy consumption in transportation sector is an important issue because it depends mainly on the use of fossil fuel.

There are many measures to reduce fuel for transportation, for example, shifting to public transportation from private cars or increasing the number of passengers per trip of each vehicle. In this study, the authors focus on the land use of urban areas. It is also effective to construct land use to reduce trip length or fuel consumption in urban areas. Some studies are carried out from the viewpoint of decreasing traffic congestion around the center of a city. Suzuki reported the effect of rearranging inhabitants over a residential district around Tokyo with respect to reducing trip time. However, the land use of the business and residential districts is fixed in this study and no optimality of land use is discussed.

The objective of this study is to achieve optimal land use in urban areas to attain minimal trip length or fuel consumption. The authors have developed two models to investigate urban land use. One minimizes the total trip length under a condition of constant congestion rate (called the minimal-trip-length model), the other minimizes the fuel consumption directly where congestion generates endogenously in the model (called the minimal-fuel model). Optimal solutions derived from numerical simulations are compared with the actual land use in the Central Business District (CBD) of Tokyo.

The following are assumed in this study.

1. The city has a shape of circle;

2. The land use is for residential and business activities as well as radial and circular roads;

3. only automobile transportation is considered;

4. commuting and business trips occur;

5. all inhabitants commute to and from business areas;

6. radial and circular roads can exist at any place;

7. trips take the shortest path; and

8. a stable state is considered.

The minimal-trip-length model neglects the time domain while the minimal-fuel model takes commuting and business trip hours into account.

(1) Minimal-trip-length model: Actual cities generally suffer traffic congestion and the construction of more roads is usually advocated to mitigate it. Since no congestion is ideal, this model describes a constant congestion city, which implies stable and smooth transportation ensues. This condition gives a proportional relationship between total fuel consumption and total trip length because the fuel consumption-rate is constant as the traffic condition is always the same at any place in the city. Therefore, the optimum land use of minimal fuel consumption can be achieved by minimization of the total trip length.

(2) Minimal-fuel model: There is trade-off relationship between trip length and fuel consumption. Trip length can decrease when business and residential areas concentrate around the center of a city, while fuel consumption may increase because traffic congestion occurs heavily around the center. This model allows traffic congestion in contrast with the minimal-trip-length model if the total fuel consumption is reduced. In other words, this model takes energy consumption as the priority, not the ideal traffic condition.

(1) Use of circular roads: It is assumed that trips occur between any places. Path connecting two points consist of the combination of radial and circular roads. The condition where circular road is used is represented by Eq. 1 from the viewpoint of the shortest path.

(1)?<2,

where ? denotes the angle between origin and destination. If the angle of two points is within 2 radian, the path does not pass through the center of the city and gets the destination point via circular road. The model assumes to have a rotary area which only road occupies at the center because it is impossible for trips to pass just on the center.

Inward direction:

Case 1 Origin O1 in the area A to the inner area of C.

Case 2 Origin O1 in the area A to the outer area of B.

Outward direction:

Case 2 (the same as above).

Case 3 Origin O2 in the area C to the outer area of A or B.

(3) Trips in circular direction: As far as circular road concerns, trips passing along the circle of radius r occur in the case where origin or destination is/are located on the circle, which Fig. 2 shows.

(4) Principle of trips: It is assumed that inhabitants at the origin (X₁) who travel to the business area (X₂) are proportional to the working population at the destination with respect to both cases of business trips and commuting trips. Eq. 2 represents this principle.

where X₁ and X₂ denote origin and destination, respectively. It is also assumed here that each person gets on one vehicle. The number of trips multiplied by generating rate give the load of traffic.

(1) Trips in radial directions: Business trips in radial directions are formulated as Eq. 3 and Eq. 4 below and based on the trip patterns shown in Fig. 1. Inward and outward direction have the same number of trips.

The first term means the trips which occur between the points on the circle of radius r and the points in outer area, and the second between the points both of which are on the circle. Here, trips passing across the angle of ? in an anti-clockwise direction occur under the following condition.

(1) Trips in radial directions: Commuting trips connecting residential and business areas can be formulated in the same way as for business trips, which Eq. 7 and Eq. 8 represent. It should be noted that the amounts of inward and outward trips are different in this case in contrast with the case of the business trips.

The left-hand side means road area supplied at radius r, while the right-hand side represents road area which is needed to deal with the traffic at that point. Eq. 10 shows that the constant-congestion condition implies that at any point in the city the road supply and demand are equal.

(1) Objective function: The main consideration of this model is the total trip length even though not all of the inhabitants commute by automobile. The length of the business trips, L₁, is expressed by Eq. 11, and that of commuting trips, L₂, can be given by an analogous procedure.

Tags: trip length

Posted in Applied Energy |

Hydrogen Gas and Hot Water Waste Heat

After the oil-supply disruptions and price shocks of 1970s, emphasis on energy saving has been intense on the industrial sector in Japan, e.g. by promoting the use of heat-cascading systems.

In transporting waste heat through pipelines, the supply of waste heat is usually located far from demand, which can make the cost of facilities (mainly pipelines) both prohibitive and unprofitable. One option is to utilize chemical reactions, which can lead to the transport of heat more intensively than via vapor or hot water in pipelines. As a result the methanol-gas decomposition/synthesis reaction, and hydrogen-absorbing alloys have been investigated as requisite technologies in Japan.

Waste heat is recovered by the endothermic reaction of methanol decomposition (CH₃O CO+2H₂). The carbon monoxide and hydrogen gases derived are then transported through the pipeline to the demand side. These gases are synthesized to methanol at high temperature and pressure (250°C, 50 atm). This synthesis reaction is exothermic, which enables heat release. The synthesized methanol is carried back through another pipeline and this closed cycle is repeated continuously.

Hydrogen-absorbing alloys can react reversibly with hydrogen gas. Heat energy is generated through hydrogen absorbing process, and the metal hydride emits hydrogen gas when heat is added to it. These characteristics can be applied to waste-heat transporting system.

Processes for transportation are as follows:

1. Waste heat is injected to Reactor 1, and the metal hydride in it emits hydrogen gas;

2. The emitted hydrogen gas is carried to Reactor 4, and the metal in it absorbs the hydrogen gas;

3. The heat generated through the exothermic reaction can be utilized at the demand side.

To provide heat energy continuously at the demand side,the metal in Reactor 3 is required to absorb hydrogen gas during the processes quoted above 1–3. So, initially the metal hydride is placed in Reactor 2, and hydrogen gas is carried to Reactor 3 when the waste heat is injected to the metal hydride of Reactor 2.

The total costs (i.e. facility costs and energy cost for running the system) of waste-heat transportation systems with different energy carriers (vapor, hot water, methanol or hydrogen gas) are evaluated here.

• The demanded temperature and pressure are set as restrictive conditions besides the calorific balance between supply and demand;

• Heat and pressure losses incurred in the pipeline transmission are included in the total cost by evaluating the cost of energy for boosting or compressing;

• The model calculates the optimal system cost by evaluating the pipeline cost under various specifications (e.g. velocity and pressure) for the use of the pipelines.

Type of heat demand: heating, cooling or hot-water supply.

Heat requirement (temperature and pressure):

heating – hot water (above 110°C) or vapor (above 1.5 atm),

cooling – hot water (above 180°C) or vapor (above 9.0 atm),

hot-water supply – hot water or vapor (above 60°C).

Type of the hydrogen absorbing alloy: LaNi5,

Time period for hydrogen absorbing=30 m, Cost of the metal 5,000 yen/kg.

Pressure loss of the transmission pipeline:

Liquid flow – subject to the Darcy – Weisbach’s law

Gas flow – subject to the equation of heat loss for isothermal fluids.

Heat loss of the transmission pipeline:

Calculated from the knowledge of the environmental temperature, heat conduction and transmission ratios.

Cost of facilities and utilities:

Heat exchanger – 10,000 yens/Mcal.h, Thermal efficiency=0.8

Refrigerator (steam absorption type) – 30,000 yens/Mcal.h, COP=1.2

Refrigerator (hot-water absorption type) – 27,000 yens/kW, COP=0.5

Compressor – 100,000 yen/kW, power efficiency=0.75

Reactor for methanol decomposing/synthesis – 24,660 yens/t(gas) · s (including compressor), thermal efficiency=0.9

Pipeline: Cost of construction (yen/km) is linearly estimated by the caliber and the distance of plumbing

Electricity – 19.69 yen/kWh Urban gas – 5.0 yen/Mcal

About modeling of heat-transport process using hydrogen absorbing alloys

Temporal change of hydrogen gas flow and heat energy supplied corresponding to the change of the pressure condition are investigated. To make the investigation easier, the hypothetical system of hydrogen transmission system shown in is assumed.

Fig. 3 shows that energy costs for the compressor are relatively high in utilizing the methanol reaction and that reactor makes the total cost rather prohibitive in utilizing hydrogen-absorbing alloys.

Fig. 4 indicates the field of demand (regarding quantity of heat energy transmitted, and distance between the heat supplier and demand point) suitable for each transportation system.

• By utilizing methanol reaction, the heat energy can be transmitted intensively, which decreases the cost of the pipelines, and makes the system more cost-effective.

• The transmission system using vapor is relatively suitable for large heat-demands located near the supplier.

It is useful to compare the cost of the transportation system with that of the case of not introducing the transportation system and obtaining the required heat by burning fossil fuels.

Tags: hydrogen gas

Posted in Applied Energy |

Heat Transfer Rate of Natural Convection

Many research papers have been published concerning steady-state heat transfers from concentrated heat sources within enclosed spaces. Significant natural-convections occur in such diverse situations as from heaters in ovens, within enclosed micro-electronic component and electrical switch-box enclosures, and during fuel-oil storage-tank heating. Studies (e.g. [10–16]) of heat transfers from horizontal cylinders, in free space or close to adjoining walls, have indicated changes in the rate of natural convection depending upon the spacings between the heaters, between each heater and the wall, and on whether the wall is vertical or horizontal.

The convective currents associated with horizontal, relatively high-temperature single cylinders or wires in free space have been studied intensively: the review by Morgan contains a comprehensive compilation of the pertinent correlations. Eckert and Soehngen considered the effect, on natural convection, of offsetting the middle horizontal cylindrical heater of a vertically-aligned array of three such heaters, each of 22.3 mm diameter (d). Lieberman and Gebhart subsequently investigated the effect of varying the spacing between thin, heated wires (each with d=0·127 mm). Sparrow and Boesneck compared the effects of a range of horizontal and vertical offsets with that for no offset using only two horizontal cylinders, each of 38 mm diameter. These studies concluded that, for the considered circumstances, the Nusselt number of the upper heater decreased to 87% of that of the lower heater, depending upon the vertical separation, because of the warm plume rising from the lower cylinder. However, if the upper heater was offset transversely by 0·5d from the vertical plane passing through the lower heater, the Nusselt number increased by 3%, because the plume from the lower heater entrains cooler air from the surroundings and it approaches the upper heater at a higher velocity.

The effect upon the rate of natural convection from a horizontal cylindrical heater (with d=6.35 mm), of the presence of two confining vertical parallel plates was studied by Marsters. Sparrow and Pfeil carried out a similar study, but employed a larger horizontal cylinder (of 38 mm diameter) situated between two walls which formed a vertical channel. Tokura et al.. examined the behaviours of 3- and 5-cylinder cylindrical arrays, with d=28.5 mm, also confined between vertical parallel plates. These studies achieved rate-of-heat transfer enhancements of up to 40%, as a result of increasing the spacing between the cylinders and the wall. With the aid of flow visualisation, Al-Alusi and Bushnell showed that a chimney effect existed to induce air flows around a 3-cylinder (each with d=25.4 mm) horizontal array situated along a single vertical wall.

Sparrow and Ansari examined the heat-transfer characteristics of three configurations, namely a heated, horizontal cylinder in close proximity to three types of walls: (i) a vertical wall situated to the side of the cylinder; (ii) a horizontal wall situated beneath the cylinder, and (iii) a corner formed by vertical and horizontal walls with the cylinder within its included angle. The geometry has the following effects on the rate of steady-state natural convection. The presence of (i) a closely-positioned side wall (in an otherwise open space) usually reduces the cylinder’s Nusselt number in relation to that occurring when no wall was present, but this value is increased if the horizontal spacing exceeds 0.25 of the cylinder’s diameter (i.e. Nu/Nu_?>1 for S_h/d>0.25); and (ii) a plane horizontal wall beneath the heater always reduces the relative value of the Nusselt number, i.e. Nu/Nu_?<1.0 throughout. When the heater was positioned near the corner, i.e. (iii) the relative rate of natural convection ( / _?) was found to be reduced to as low as 0.6.

Extensive experimental and numerical studies of the heat transfers across various fluid-filled cylindrical annuli exist. Kuehn and Goldstein presented an experimental study, which analysed the effects of annular eccentricity on natural-convection heat transfer through air. The inner object was a relatively-large, 35.6 mm-diameter cylindrical heater in a 92.5 mm-diameter cylindrical enclosure: a non-linear relationship was established between the convective-heat transfer and the heater elevation. Cho et al. also derived numerically corresponing non-linear relationships when the inner cylinder was traversed both horizontally and vertically. Shilston carried out an experimental investigation by traversing a 28 mm-diameter cylindrical heater along the vertical line of symmetry in a 100 mm×100 mm square-sectioned, 600 mm long enclosure. With the aid of a Mach–Zehnder interferometer, the rate of natural convection heat transfer was found to decrease as the heater was moved away from a concentric position towards the upper horizontal wall, although enhancement was noted when it was located just below the central position.

In the experimental studies of Warrington and Powe, particular emphasis was placed on how the rates of heat transfer from a variety of hotter, axi-symmetrical inner bodies (i.e. sphere, cube and cylinder) are influenced by their locations in a cubical (267 mm×267 mm×267 mm) enclosure. When compared with the studies of Weber et al. involving a spherical enclosure, it was revealed that the enclosure size and the gap between the spherical inner body and the enclosure had more significant influences on the rate of natural convection than the shape of the enclosure. The numerical investigation of Deschamps and Desrayaud concerned a cylindrical or line-heat source, situated in a rectangular-sectioned enclosure. At higher values of Ra, there was increased convective activity near the upper horizontal wall (i.e. the ceiling). The experimental and numerical study of Zhao et al., for three uniform-heat-flux emitting cylinders, placed side-by-side near the base of a rectangular enclosure (which served as the heat sink), revealed that the centrally-positioned cylinder was warmest and experienced the lowest mean convective heat-transfer because of the restricted peripheral flows around it.

While many research studies have considered symmetrically-located heaters in enclosures, practical systems, such as encountered in ovens, often require the installation of cylindrical heaters along the enclosure walls. There is a wide variation of the Nusselt number values depending upon the heater location, and even then, there are insufficient experimental data for the intricate and complicated air-flow processes. Thus an attempt is made, in this investigation, to quantify the dependence of the steady-state rate of natural convection from a heater with its location in a rectangular enclosure.

This consisted of a horizontal cylindrical heater located within a metal box with glazed end-plates. This square-sectioned (350 mm×350 mm) enclosure, 750 mm long horizontally, was fabricated from 1.2 mm thick mild-steel sheet, and its inner surfaces finished with a high-temperature matt-black paint (=0.85±0.01). High–temperature glass plates were used to cover the ends of the enclosure. The enclosure was the cold surface, being cooled by the surrounding ambient air, i.e. similar to the experimental method of Zhao et al. The lower horizontal surface will be referred to as the base while the upper horizontal surface is termed the crown.

The first heat-source to be employed in the enclosure was a single cylindrical electric heater, of 9.5 mm diameter and an overall heated length of 590 mm. The maximum power-input flux for the electric element was limited to 990 W/m² of heating surface, so resulting in relatively low heater-temperatures (?363 K), and hence the radiation contribution did not dominate the heat transfers. An automatic a.c. voltage stabiliser and an auto-transformer were employed to provide a stable power supply to maintain steady-state temperature-distributions. The power consumption of the heater was measured with a calibrated wattmeter.

Tags: natural convection

Posted in Applied Energy |