CHAPTER 1 INTRODUCTION
The present energy scenario has stimulated active research interest in nonpetroleum, renewable, and non-polluting fuels. Increased dependence on imported fuel and regulations on the exhaust emissions of vehicles established by the Environmental Protection Agency (EPA) has increased the need for alternative fuels. Alternative fuels from agricultural crops are the desirable substitute for diesel because they are sustainable and have fewer emissions than petroleum products.But some of the alternative fuel does not have good properties, the properties related to fluidity and their viscosity. The proper atomisation of the fuel used is of great importance. Power
Air After Cooler
Friction and Radiation
Liquid Jacket Water
Exhaust
30% 38%
18%
7% 7%
Figure 1.1 Energy Distributions from Engine But it needs greater energy requirements to do soand thus increase the cost of vehicle. it is evident from the diagram that Contemporary car engines exchange app. 30-40% of heat generated in the process of fuel combustion into useful mechanical work, The remaining heat is emitted to the environment through the exhaust gases and the engine cooling systems. The single largest amount of unused heat from the engine is the exhaust heat, which contains about 30% of the fuel energy. it is evident that exhaust gases comes out from the exhaust port Page | 1
at a very high temperature, it has been seen that in diesel engines exhaust emission are at the temperature of 250deg centigrade to 450deg centigrade (approx.) this high temperature might be used in an efficient way. If the waste heat were put to appropriateuse, there would be great energy or fuel savings. The definition of waste heat includes the following: unburned but combustible fuel; sensible and latent enthalpy discharge from exhaust gas mixtures; and sensible heat discharge in liquid waste (Al Rabghi et al 1993) . But the energy lost along with the exhaust gas, can be ascribed to the inefficient utilisation of the available energy, which makes research into the recovery of waste energy and subsequent efficiency and fuel utilization improvement, a viable option. To recover this form of energy some type of device is required which can efficiently use the energy coming out with the exhaust gases, so the idea of the use of heat exchanger came in to existence, which can efficiently transfers the heat from exhaust gases to the other fluid. a common shell and tube heat exchanger can be used for this purpose, and the designed heat exchanger might be said exhaust gas heat exchanger, this heat exchanger should be capable in easy maintenancehigh heat transfer rates and low fouling rates also. This report explains the use of exhaust gases through heat exchanger, the heat of exhaust gases might be used in heating an alternative fuel.the alternative fuel taken is biodiesel,which have higher calorific value but less than diesel fuel, inspite of this it is widely accepted as a alternative fuel.Now the question arises why? Bio-diesel is heated, basically there is some problems with the direct use of bio-diesel, the high viscosity and density of biofuels leaded to problems in the injection system and combustion chamber of the diesel engines for a long- term usage. The high viscosity problem of bio-fuel has been solved inmany ways one of them is given below.
1.1
Solution to Improve the Properties of Vegetable Oil
1.1.1 Preheating of Vegetable oil Preheating of vegetable oil means heat the oil before the actual heating takes place inside the cylinder. This is done because when the cold vegetable oil gets enter to the cylinder its initial temperature is not so high. So at the time of combustion proper atomization of the Page | 2
fuel cannot take place and sufficient air –fuel mixture is not burn inside the cylinder, it can cause incomplete combustion and will reduce the life of the engine. So Preheating is done to overcome this difficulty and also it will improve the quality of vegetable oil.
1.2 Summary From the above points it is found that bio-diesel is the fuel rapidly growing in use, and if we use it in diesel engines then it should have good fluidity, low viscosity and good atomization which can only possible if we heat it either externally or with help of an alternativearrangement, this alternative arrangement might be high temperature exhaust gases coming out of the exhaust port, these exhaust gases if let free in to atmosphere can cause great trouble related to health. It is also a fact that use of some electrical mean to heat the fuel externally can cause extra cost of operating.so a idea extract from above is to design such a device which can eliminate the use of this external power andcan also reduce the temperature of the exhaust gases to a greater extent.
Page | 3
CHAPTER 2 LITERATURE REVIEW
Following is the detail that gives the brief idea on heat exchanger design and waste heat recovery through the heat exchanger and waste heat use in different applications till date. 1. Automobile exhaust gas of using heat pipe heat exchangers for heating applying automotive exhaust gas is studied and the calculation method is developed. Practical heat pipe heat exchanger is set up for heating HS663, a large bus. Simple experiments are carried out to examine the performance of the heat exchanger. It is shown that the experimental results, which indicate the benefit of exhaust gas heating, are in good agreement with numerical results.
Figure 2.1 Schematic Diagram of Heating System 2.
Design of a heat exchanger to reduce the exhaust temperature in a spark-ignition engine. A design of experiments (DOE) technique was used to design an exhaust heat exchanger to reduce the exhaust gas temperature under high load conditions in a spark-ignition engine. Through a limited number of experiments, the DOE evaluated the influence and the interaction of eight selected design parameters of the heat exchanger that affect the cooling performance of the exhaust gas. The heat exchanger was installed between the Page | 4
exhaust manifold and the inlet of the close-coupled catalytic converter (CCC) to avoid thermal aging. To maximize the heat transfer between the exhaust gas and coolant, fins were implemented at the inner surface of the heat exchanger. The design parameters consisted of the fin geometry (i.e., length, thickness, arrangement, and number of fins), coolant direction, exchanger wall thickness, and the length of the heat exchanger. The DOE results were analyzed and the acceptable range of each design parameter is discussed.
Figure 2.2 Engine test facility and coolant circulation scheme 3.
Effect of cooling the recirculated exhaust gases on diesel engine emissions. Although
combustion is essential in most energy generation processes, it is one of the major causes of air pollution. Spiral fin exhaust pipes were designed to study the effect of cooling the recirculated exhaust gases (EGR) of Diesel engines on the chemical composition of the exhaust gases and the reduction in the percentages of pollutant emissions. The gases examined in this study were oxides of nitrogen (NOx), carbon dioxide (CO2) and carbon monoxide (CO). In addition, O2 concentration in the exhaust was measured. The two designs adopted in this study were exhaust pipes with solid and hollow fins around them. The first type uses air flow around the fins to cool the exhaust gases. The second type consists of hollow fins around the exhaust pipe to allow cooling water to flow in the hollow passage. Different combinations and arrangements of the solid and hollow fins exhaust pipes were used. It was found that decreasing the temperature of the Page | 5
EGR resulted in reductions in the oxides of nitrogen (NOx) and carbon dioxide (CO2) but increased the carbon monoxide (CO) in the exhaust gases. In addition, the oxygen (O2) concentration in the exhaust was decreased. As a general trend, the percentages of reduction in the NOx gas concentrations were lower than the percentages of increase in the CO emissions as a result of cooling the EGR of a Diesel engine by a heat exchanger. Using water as a cooling medium decreased the exhaust gases temperature and the amount of pollutants more than did air as a cooling medium. In a separate series of tests, increasing the cooled EGR ratios decreased the exhaust NOx but increased the particulate matter concentrations in the exhaust gases. 4.
Performance of shell and dimpled tube heat exchangers for waste heat of diesel engine recovery
by
shell
and
tube heat
exchangers.
The exchanger
heat duty,
overall heat transfer coefficient, effectiveness and tube side friction factor are investigated as functions of the tube surface geometry (plain or dimpled), the flow type (counter or parallel), the tube Reynolds number and the shell side heat capacity rate. Water and the exhaust gases of a Diesel engine are passed inside the tube and the shell, respectively. The heat transfer characteristics increase with an increase in tube Reynolds number and the shell side heat capacity rate, for all the flow types and the surface geometries examined. The counter-flow, shell-and-dimpled-tube heat exchanger, compared with that exchanger having a plain tube, increases the heat duty and the overall heat transfer coefficient by 80%, and the heat exchanger effectiveness increases by 35%. For the parallel-flow,
shell
and
dimpled
tube heat
exchanger,
the heat duty,
the
overall heat transfer coefficient and the effectiveness increase by 30, 55, and 25%, respectively. At the same time the dimpled tube increases the tube side friction factor by 600% over that of the plain tube. The rate of waste heat recovered from the exhaust gases of the Diesel engine by the counter-flow, shell-and-dimpled-tube heat exchanger is equal to 10% of the maximum brake power of the engine running at 1500 rpm, and the tube Reynolds number equal to 8875. 5.
Stainless steel finned tube heat exchanger design for waste heat recovery associated with hydrocarbon combustion in the transport industry is identified as a significantly underutilised energy resource. The aim of this project was to investigate the recovery of waste Page | 6
heat in a small scale system for the purpose of electrical conversion in order to serve as a secondary energy source. A theoretical analysis concerning the design and construction of the system, utilising researched theory and a control volume based simulation program of the recovery system, is presented. It was found that heat exchangers for the required duty are not readily available in South Africa. A high pressure, cross flow, stainless steel finned tube heat exchanger with a water side pressure rating of 2 MPa was therefore designed and constructed. By using the exhaust gases of a continuous combustion unit as an energy source and water as the working fluid, efficiencies of up to 74% in direct steam generation testing were obtained. 6.
Diesel engine and improving fuel consumption are important in meeting government regulations and society needs. Use of the Cooled Exhaust Gas Recirculation (EGR) system is one of the most effective techniques currently available for reducing NOx and PM emissions. However, the EGR system has a trade-off between NOx and PM emissions at high loads. In the present study, engine dynamometer experiments have been performed to investigate the heat exchange effectiveness of EGR coolers with shell & tube-type and stack-type. The results show that the heat transfer effectiveness of the stack-type EGR cooler is 25-50 % higher than that of the shell & tube type due to an increased surface area and a better mixing of the exhaust gas flow.
Figure 2.3 Schematic diagram for the EGR cooler performance test Page | 7
Shell & Tube TypeStack Type
Figure 2.4 Front-view photographs of the two EGR coolers. 7.
A waste heat recovery system composed of a two phase cooling system, an exhaust heat exchanger, and mini-turbine (expander) has been proposed by Henry Works, Inc to generate auxiliary power via harvesting engine cooling and exhaust heat loss from heavy duty vehicles. The objective of this research is to evaluate the two phase cooling system through engine dynamometer testing and obtain initial test data for the development of the waste heat recovery system. Engine dynamometer experimentation for evaluating two phase cooling has been conducted using a Perkins diesel engine. During the two phases cooling phase, the coolant temperature showed less than 1 °C variation in the cooling path and the cylinder head temperature was more uniform than that of single phase cooling. As the saturated vapour pressure increases during two phase cooling, the cylinder head and coolant temperatures also increase. Thus, the maximum pressure of the saturated vapour in the two phases cooling is limited by allowable cylinder head temperature that is determined by cylinder head distortion, abnormal combustion, exhaust emissions, etc. The water coolant mixed with trifluoroethanol showed lower cylinder head temperature than pure water coolant at higher vapour pressure of the coolant. Based on the measured values in the engine dynamometer experiment, the potential power output of the proposed waste heat recovery system under the same engine operating conditions in this study ranges from 0.47 KW ~ 1.05 KW.
Page | 8
8.
Automotive engines reject a considerable amount of energy to the ambience through the exhaust gas. Significant reduction of engine fuel consumption could be attained by Recovering of exhaust heat by using thermoelectric generators. One of the most important issues is to develop an efficient heat exchanger which provides optimal recovery of heat from exhaust gases. The work presents a design and performance measurements of a prototype thermoelectric generator mounted on self-ignition (Diesel) engine. Using theprototype generator as a tool, benchmark studies were performed for improvements in the heat exchanger including determination of temperature distribution and heat flux density.
Page | 9
CHAPTER 3 DESIGN OF EXHAUST GAS HEAT EXCHANGER
3.1
Selection of Heat Exchanger A variety of heat exchanger is available, and the question becomes which are to
choose for a given application. In addition, for each type ( core construction), either a large type of geometrical variable (such as those associated with each component of the shell and tube exchanger) or a large number of surface geometries( such as those for plate, extended surface, or regenerative exchanger) are available for selection for heat exchanger various type of heat exchanger description is given . 3.1.1 Double-Pipe Heat Exchangers This exchanger usually consists of two concentric pipes with the inner pipe plain or finned, as shown in Fig. 3.1 One fluid flows in the inner pipe and the other fluid flows in the annulus between pipes in a counterflow direction for the ideal highest performance for the given surface area. However, if the application requires an almost constant wall temperature, the fluids may flow in a parallel flow direction. This is perhaps the simplest heat exchanger.
Figure 3.1 Double Pipe Heat Exchangers
Page | 10
FlowDistribution is no problem, and cleaning is done very easily by disassembly. This configuration is also suitable where one or both of the fluids is at very high pressure,because containment in the small-diameter pipe or tubing is less costly than containment in a largediameter. Double-pipe exchangers are generally used for small-capacityapplications where the total heat transfer surface area required is 50m2 (500 ft2) or less because it is expensive on a cost per unit surface area basis .But it can create problem at high temperatures and also when the fluid used is of high viscosity may lead to corrosion and choking. 3.1.2 Spiral Tube Heat Exchangers These consist of one or more spirally wound coils fitted in a shell. Heat transfer rate associated with a spiral tube is higher than that for a straight tube. In addition, a considerable amount of surface can be accommodated in a given space by spiralling. Thermal expansion is no problem, but cleaning is almost impossible. 3.1.3 Plate-Type Heat Exchangers Plate-type heat exchangers are usually built of thin plates (all prime surfaces). The platesare either smooth or have some form of corrugation, and they are either flat or wound in an exchanger. Generally, these exchangers cannot accommodate very high pressures, temperatures, or pressure and temperature differences. Plate heat exchangers (PHEs) {can be classified as gasketed, welded (one or both fluid passages), or brazed, depending on the leak tightness required. Other plate-type exchangers are spiral plate, lamella, and plate coil exchangers.
Figure 3.2 Plate Type Heat Exchanger Page | 11
Advantages and Limitations, Some advantages of plate heat exchangers are as follows. They can easily be taken apart into their individual components for cleaning, inspection, and maintenance. The heat transfer surface area can readily be changed or rearranged for a different task or for anticipated changing loads, through the flexibility of plate size, corrugation patterns, and pass arrangements. High shear rates and shear stresses, secondary flow, high turbulence, and mixing due to plate corrugation patterns reduce fouling to about 10 to 25% of that of a shell-and-tube exchanger, and enhance heat transfer. Very high heat transfer coefficients are achieved due to the breakup and reattachment of boundary layers, swirl or vortex flow generation, and small hydraulic diameter flow passages. Because of high heat transfer coefficients, reduced fouling, the absence of bypass and leakage streams, and pure counterflow arrangements, the surface area required for a plate exchanger is one-half to one-third that of a shell-and tube exchanger for a given heat duty, thus reducing the cost, overall volume, and space requirement for the exchanger. Also, the gross weight of a plate exchanger is about onesixth that of an equivalent shell-and-tube exchanger. Leakage from one fluid to the other cannot take place unless a plate develops a hole. Since the gasket is between the plates, any leakage from the gaskets is to the outside of the exchanger. The residence time (time to travel from the inlet to the outlet of the exchanger) for different fluid particles or flow paths on a given side are approximately the same. There are no significant hot or cold spots in the exchanger that could lead to the deterioration of heat-sensitive fluids. The volume of fluid held up in the exchanger is small; this feature is important with expensive fluids, for faster transient response, and for better process control. Finally, high thermal performance can be achieved in plate exchangers. The high degree of counter flow in PHEs makes temperature approaches of up to 18C (28F) possible. The high thermal effectiveness (up to about 93%) facilitates economical low-grade heat recovery. The flow-induced vibrations, noise, thermal stresses, and entry impingement problems of shell-and-tube exchangers do not exist for plate heat exchangers. Some inherent limitations of the plate heat exchangers are caused by plates and gaskets as follows. The plate exchanger is capable of handling up to a maximum pressure of about 3 MPa gauge (435 psig) but is usually operated below 1.0 MPa gauge (150 psig). The gasket materials (except for the PTFE-coated type) restrict the use of PHEs in highly corrosive applications; they also limit the maximum operating temperature to 260degC Page | 12
(500degF) but are usually operated below 150degC (300degF) to avoid use of expensive gasket materials. Gasket life is sometimes limited. Frequent gasket replacement may be needed in some applications. Pinhole leaks are hard to detect. For equivalent flow velocities, pressure drop in a plate exchanger is very high compared to that of a shell-andtube exchanger. However, the flow velocities are usually low and plate lengths are „„short,‟‟so the resulting pressure drops are generally acceptable. For some cases, heat exchanger duties with widely different fluid flow rates and depending on the allowed pressure drops of the two fluids, an arrangement of a different number of passes for the two fluids may make a PHE advantageous. However, care must be exercised to take full advantage of available pressure drop while multipassing one or both fluids. Because of the long gasket periphery, PHEs are not suited for high-vacuum applications. PHEs are not suitable for erosive duties or for fluids containing fibrous materials. In certain cases, suspensions can be handled; but to avoid clogging, the largest suspended particle should be at most one-third the size of the average channel gap. Viscous fluids can be handled, but extremely viscous fluids lead to flow maldistribution problems, especially on cooling. Plate exchangers should not be used for toxic fluids, due to potential gasket leakage. Major Applications Plate heat exchangers are not well suited for lower-density gas-togas applications. They are used for condensation or evaporation of non-low-vapour densities. 3.1.4 Spiral Plate Heat Exchangers A spiral plate heat exchanger consists of two relatively long strips of sheet metal, normally provided with welded studs for plate spacing, wrapped helically around a split mandrel to form a pair of spiral channelsfor two fluids, as shown in Fig. Thus, each fluid has a long single passage arranged in a compact package. To complete the exchanger, covers are fitted at each end. Any metal that can be cold-formed and welded can be used for this exchanger. Common metals used are carbon steel and stainless steel. Other metals include titanium, Hastelloy, Incoloy, and high-nickel alloys. A spiral plate exchanger has a relatively large diameter because of the spiral turns. The largest exchanger has a maximum surface area of about 500m2 (5400 ft2) for a maximum shell diameter of 1.8m (72 in.). The heat transfer coefficients are not as high as in a plate exchanger if the plates are not corrugated. However, the heat transfer coefficient is higher than that for a shell-and-tube exchanger because of the Page | 13
curved rectangular passages. Hence, the surface area requirement is about 20% lower than that for a shell-and-tube unit for the same heat duty. The counter flow spiral unit is used for liquid–liquid, condensing, or gas cooling applications. When there is a pressure drop constraint on one side, such as with gas flows or with high liquid flows, cross flow (straight flow) is used on that side. Horizontal units are used when high concentrations of solids exist in the fluid.
Figure 3.3 Spiral Plate Heat Exchangers Tube exchanger because of the curved rectangular passages. Hence, the surface area requirement is about 20% lower than that for a shell-and-tube unit for the same heat duty. The counterflow spiral unit is used for liquid–liquid, condensing, or gas cooling applications. When there is a pressure drop constraint on one side, such as with gas flows or with high liquid flows, cross flow (straight flow) is used on that side. Horizontal units are used when high concentrations of solids exist in the fluid. The advantages of this exchanger are as follows: It can handle viscous, fouling liquids and slurries more readily because of a single passage. If the passage starts fouling, the localized velocity in the passage increases. The fouling rate then decreases with increased fluid velocity. The fouling rate is very low compared to that of a shell-and-tube unit. It ismore amenable to chemical, flush, and reversing fluid cleaning techniques because of the single passage. Mechanical cleaning is also Page | 14
possible with removal of the end covers. Thus, maintenance is less than with a shell-and-tube unit. No insulation is used outside the exchanger because of the cold fluid flowing in the outermost passage, resulting in negligibleheat loss, if any, due to its inlet temperature closer to surrounding temperature. The internal void volume is lower (less than 60%) than in a shelland-tube exchanger, and thus it is a relatively compact unit. The disadvantages of this exchanger are as follows: As noted above, the maximum size is limited. The maximum operating pressure ranges from 0.6 to 2.5MPa gauge (90 to 370 psig) for large units. The maximum operating temperature is limited to 500degC (930F) with compressed asbestos gaskets, but most are designed to operate at 200C (392F). Field repair is difficult due to construction features. This exchanger is well suited as a condenser or reboiler. It is used in the cellulose industry for cleaning relief vapors in sulfate and sulfite mills, and is also used as a thermo siphon or kettle reboiler. It is preferred especially for applications having very viscous liquids, dense slurries, digested sewage sludge, and contaminated industrial effluents. 3.1.5 Extended Surface Heat Exchangers The plate-type exchangers described previously are all prime surface heatexchangers, except for a shell-and-tube exchanger with low finned tubing. Their heat exchanger effectiveness is usually 60% or below, and the heat transfer surface area density is usually less than 700m2/m3 (213 ft2/ft3). In some applications, much higher (up to about 98%) exchanger effectiveness is essential, and thebox volume and mass are limited so that a much more compact surface is mandated. Also, in a heat exchanger with gases or some liquids, the heat transfer coefficient is quite low on one or both fluid sides. This results in a large heat transfer surface area requirement. One of the most common methods to increase the surface area and exchanger compactness is to add the extended surface (fins) and use fins with the fin density ( fin frequency, fins/m or fins/in.) as high as possible on one or both fluid sides, depending on the design requirement. Addition of fins can increase the surface area by 5 to 12 times the primary surface area in general, depending on the design. The resulting exchanger isreferred to as an extended surface exchanger. The heat transfer coefficient on extended surfaces may be higher or lower than that on unfinned surfaces. For example, interrupted (strip, louver, etc.) fins provide both an Page | 15
increased area and increased heat transfer coefficient,while internal fins in tube increase the tube-side surface area but may result in a slight reduction in the heat transfer coefficient, depending on the fin spacing. Generally, increasing the fin density reduces the heat transfer coefficient associated with fins. Flow interruptions (as in offset strip fins, louvered fins, etc.) may increase the heat transfercoefficient two to four times that for the corresponding plain (uncut) fin surface. Plate-fin and tube-fin geometries are the two most common types of extended surface heat exchangers.
Figure 3.4 Corrugated Fin Geometries for Plate-Fin Heat Exchangers: (a) Plain triangular fin; (b) plain rectangular fin; (c) Wavy fin; (d) offset strip fin; (e) Multilouver fin; (f ) Perforated fin 3.1.6 Tube-Fin Heat Exchangers These exchangers may be classified as conventional and specialized tube-fin exchangers. In a conventional tube-fin exchanger, heat transfer between the two fluids takes place by conduction through the tube wall. However, in a heat pipe exchanger (a specialized type of tube-fin exchanger), tubes with both ends closed act as a separating wall, and heat transfer between the two fluids takes place through this „„separating wall‟‟ (heat pipe) by conduction, and evaporation and
Page | 16
condensation of the heat pipe fluid. Let us first describe conventional tube-finexchangers and then heat pipe exchangers
Figure 3.5 Tube-Fin Heat Exchangers 3.1.7 Heat Pipe Heat Exchangers This type of exchanger is similar to a tube-fin exchangerwith individually finned tubes or flat (continuous) fins and tubes.
Figure 3.6 Heat Pipe Heat Exchangers Page | 17
The inner surfaces of a heat pipe are usually lined with a capillary wick (a porous lining, screen, or internally grooved wall). The wick is what makes the heat pipe unique; it forces condensate to return to the evaporator by the action of capillary force, a properly designed heat pipe can transport the energy of phase change continuously from the evaporator to the condenser. Without drying out the wick.As long as there is a temperature difference between the hot and cold gases in a heat pipe heat exchanger, Generally, there is a small temperature difference between the evaporator and condensersection [about 58C (98F) or so, and hence the overall thermal resistance of a heat pipe in a heat pipe exchanger is small. Although water is a common heat pipe fluid, other fluids are also used, depending on the operating temperature range a heat pipe in a HPHE does not have the usual adiabatic section that most heat pipeshave. Unit size varies with airflow. Small units have a face size of 0.6m (length) by 0.3m (height), and the largest units may have a face size up to 5m 3 m. In the case of gas-to liquid heat exchangers, the gas section remains the same, but because of the higher external heat transfer coefficient on the liquid side, it need not be finned externally or can even be shorter in length. The heat pipe performance is influenced by the angle of orientation, since gravity plays an important role in aiding or resisting the capillary flow of the condensate. Because of this sensitivity, tilting the exchanger may control the pumping power and ultimately the heat transfer. This feature can be used to regulate the performance of a Heat pipe heat exchangers are generally used in gas-to-gas heat transfer applications. They are used primarily in many industrial and consumer product–oriented waste heat recovery applications. 3.1.8 Tubular Heat Exchangers Tubular exchangers are widely used in industry for the following reasons. They are custom designed for virtually any capacity and operating conditions, such as from high vacuum to ultrahigh pressure [over 100 MPa (15,000 psig)], from cryogenics to high temperatures [about 1100degC (2000degF)] and any temperature and pressure differences between the fluids, limited only by the materials of construction. They can be designed for special operating conditions: vibration, heavy fouling, highly viscous fluids, erosion, corrosion, toxicity, radioactivity, multicomponent mixtures, and so on. They are the most versatile exchangers, made from a variety of metal and non-metal materials (such as graphite, Page | 18
glass, and Teflon) and range in size from small [0.1m2 (1 ft2)] to supergiant [over 105m2 (106 ft2)] surface area. They are used extensively as process heat exchangers in the petroleum-refining and chemical industries; as steam generators, condensers, boiler feed water heaters and oil coolers in power plants; as condensers and evaporators in some airconditioning and refrigeration applications; in waste heat recovery applications with heat recovery from liquids and condensing fluids; and in environmental control. 3.1.9 Shell-and-Tube Exchangers This exchanger is generally built of a bundle of round tubes mounted in a cylindrical shell with the tube axis parallelto that of the shell. One fluid flows inside the tubes, the other flows across and along the tubes. The major components of this exchanger are tubes (or tube bundle), shell, frontend head, rear-end head, baffles, and tube sheets. A variety of different internal constructions are used in shell-and-tube exchangers, depending on the desired heat transfer and pressure drop performance and the methods
employed to reduce thermal
stresses, to prevent leakages, to provide for ease of cleaning, to contain operating pressures and temperatures, to control corrosion, to accommodatehighly asymmetric flows, and so on. Shell-and-tube exchangers are classified and constructed in accordance with the widely used TEMA (Tubular Exchanger Manufacturers Association) standards (TEMA, 1999), DIN and other standards in Europe and elsewhere, and ASME (American Society of Mechanical Engineers) boiler and pressure vessel codes.
3.2
Selection From the above text it would be better to use shell and tube type of heat exchanger as
it can withstand with the high temperature and pressure of exhaustgases, and it can also fabricated in compact form from the point of view to install it at the very small space available at the exhaust pipe. It is easy to maintain and sensitive to the heat transfer rates. Another point of view of selecting shell and tube heat exchanger its easyness to design, there are various methods available for the effective design of heat exchanger.There are various components in a shell and tube heat exchanger which should be consider while designing. These components are defined as follows:
Page | 19
3.3
Heat Exchanger Component Tubes are drawn, extruded, or welded, and they are made from metals, plastics, and
ceramics, depending on the applications.
3.3.1 Shells The shell is a container for the shell fluid .Usually; it is cylindrical in shape with a circular cross section, although shells of different shapes are used in specific applications. The shell is made from a circular pipe if the shell diameter is less than about 0.6m (2 ft) and is made from a metal plate rolled and welded longitudinally for shell diameters greater than 0.6m (2 ft).. Although the tubes may have single or multiple passes, there is one pass on the shell side. To increase the mean temperature difference and hence exchanger effectiveness, a pure counter flow arrangement is desirable for a two-tube-pass exchanger. This is achieved by use of a shell having a longitudinal baffle and resulting in two shell passes. Split- and divided-flow shells are used for specific applications, such as thermo siphon boiler, condenser, and shell-side low pressure drops. 3.3.2 Nozzles The entrance and exit ports for the shell and tube fluids, referred to as nozzles are pipes of constant cross section welded to the shell and channels. They are used to distribute or collect the fluid uniformly on the shell and tube sides. the nozzles used in heat exchanger are differ from the nozzle used as a fluid metering device or in jet engines, which has a variable flow area along the flow length. 3.3.3 Front- and Rear-End Heads These are used for entrance and exit of the tube fluid; in many rear-end heads, a provision has been made to take care of tube thermal expansion. The front-end head is stationary, while the rear-end head could be either stationary (allowing for no tube thermal expansion) or floating, depending on the thermal stresses between the tubes and shell. The major criteria for selection of the front-end head are cost, maintenance and inspection, hazard due to mixing of shell and tube fluids, and leakage to ambient and operating pressures. The major criteria for selection of the rear-end head are the allowance for thermal stresses, a Page | 20
provision to remove the tube bundle for cleaning the shell side, prevention of mixing of tube and shell fluids, and sealing any leakage path for the shell fluid to ambient. 3.3.4 Baffles Baffles may be classified as transverse and longitudinal types. The purpose of longitudinal baffles is to control the overall flow direction of the shell fluid such that a desired overall flow arrangement of the two fluid streams is achieved. Transverse baffles may be classified as plate baffles and grid (rod, strip, and other axial-flow) baffles. Plate baffles are used to support the tubes during assembly and operation and to direct the fluid in the tube bundle approximately at right angles to the tubes to achieve higher heat transfer coefficients. Plate baffles increase the turbulence of the shell fluid and minimize tube-to-tube temperature differences and thermal stresses due to the crossflow. Single- and doublesegmental baffles are used most frequently due to their ability to assist maximum heat transfer (due to a high-shell-side heat transfer coefficient) for a given pressure drop in a minimum amount of space. Triple and no-tubes-in-window segmental baffles are used for low-pressure-drop applications. The choice of baffle type, spacing, and cut is determined largely by flow rate, desired heat transfer rate, allowable pressure drop, Rod (or bar) baffles, the most common type of grid baffle, used to support the tubes and increase the turbulence of the shell fluid. 3.3.5 Tube sheets These are used to hold tubes at the ends. A tubesheet is generally a round metal plate with holes drilled through for the desired tube pattern, holes for the tie rods (which are used to space and hold plate baffles), grooves for the gaskets, and bolt holes for flanging to the shell and channel. To prevent leakage of the shell fluid at the tubesheet through a clearance between the tube hole and tube, the tube-to-tubesheet joints are made by many methods, such as expanding the tubes, rolling the tubes,hydraulic expansion of tubes, explosive welding of tubes, stuffing of the joints, or welding or brazing of tubes to the tube sheet. The leak-free tube-to-tube sheet jointmade by the conventional rolling process. 3.3.6 Tubes Round tubes in various shapes are used in shell-and-tube exchangers. Some of the enhanced tube geometries used in shell-and-tube exchangers areSerpentine, helicaland Page | 21
bayonet are other tube shapes that are used in shell-and-tube exchangers. In most applications, tubes have single walls, but when working with radioactive.
3. 4
Design Consideration of Heat Exchanger
3.4.1 Fluid allocation: shell or tubes Where no phase change occurs, the following factors will determine the allocation of the fluid streams to the shell or tubes. 3.4.1.1 Corrosion The more corrosive fluid should be allocated to the tube-side. This will reduce the cost of expensive alloy or clad components. 3.4.1.2 Fouling The fluid that has the greatest tendency to foul the heat-transfer surfaces should be placed in the tubes. This will give better control over the design fluid velocity, and the higher allowable velocity in the tubes will reduce fouling. Also, the tubes will be easier to clean. 3.4.1.3 Fluid Temperatures If the temperatures are high enough to require the use of special alloys placing the higher temperature fluid in the tubes will reduce the overall cost. At moderate temperatures, placing the hotter fluid in the tubes will reduce the shell surface temperatures, and hence the need for lagging to reduce heat loss, or for safety reasons. 3.4.1.4 Operating Pressures The higher pressure stream should be allocated to the tube-side. High-pressure tubes will be cheaper than a high-pressure shell. 3.4.1.5 Pressure Drop For the same pressure drop, higher heat-transfer Coefficients‟ will be obtained on the tube-side than the shell-side, and fluid with the lowest allowable pressure drop should be allocated to the tube-side. Page | 22
3.4.1.6Viscosity Generally, a higher heat-transfer coefficient will be obtained by allocating the more viscous material to the shell-side, providing the flow is turbulent. The critical Reynolds number for turbulent flow in the shell is in the region of 200. If turbulent flow cannot be achieved in the shell it is better to place the fluid in the tubes, as the tube-side heat transfer coefficient can be predicted with more certainty. 3.4.1.7 Stream flow-rates Allocating the fluids with the lowest flow-rate to the shell-side will normally give the most economical design.
3.4.2 Shell and tube Fluid velocities High velocities will give high heat-transfer coefficients but also a high-pressure drop. The velocity must be high enough to prevent any suspended solids settling, but not so high as to cause erosion. High velocities will reduce fouling. Plastic inserts are sometimes used to reduce erosion at the tube inlet. Typical design velocities are given below:
3.4.2.1 Liquids Tube-side, process fluids: 1 to 2.5 m/s, maximum 4 m/s if required to reduce fouling; water: 1.5 to 2.5 m/s. Shell-side: 0.3 to 1 m/s. 3.4.2.2 Vapours For vapours, the velocity used will depend on the operating pressure and fluid density; thelower values in the ranges given below will apply to high molecular weight materials. Vacuum 50 to 70 m/s Atmospheric pressure 10 to 30 m/s High-pressure 5to10m/s 3.4.2.3 Pressure drop In many applications the pressure drop available to drive the fluids through the exchanger will be set by the process conditions, and the available pressure drop will vary from a few millibars in vacuum service to several bars in pressure systems. When the Page | 23
designer is free to select the pressure drop an economic analysis can be made to determine the exchanger design which gives the lowest operating costs, taking into consideration both capital and pumping costs. However, a full economic analysis will only be justified for very large, expensive, exchangers. The values suggested below can be used as a general guide, and will normally give designs that are near the optimum.
3.4.2.4 Liquids Viscosity <1mNs/m2
35 kN/m2
1to10mNs/m2 50 70 kN/m2 3.4.2.5 Gas and vapours High vacuum
= 0.4 -0.8 kN/m2
Medium vacuum
= 0.1 × absolute pressure
1to2bar
= 0.5 × system gauge pressure
Above 10 bar
= 0.1× system gauge pressure
When a high-pressure drop is utilised, care must be taken to ensure that the resulting high fluid velocity does not cause erosion or flow-induced tube vibration.
3.5
Method Employed For Design The thermal design of heat exchangers is directed to calculate an adequate surface
area to handle the thermal duty for the given specifications whereas the hydraulic analysis determines the pressure drop of the fluids flowing in the system, and consequently the pumping power or fan work input necessary to maintain the flow. The most common problems in heat exchanger design are rating and sizing. The rating problem is evaluating the thermo-hydraulic performance of a fully specified exchanger. The rating program determines the heat transfer rate and
the fluid outlet temperatures
for prescribed fluid flow rates, inlet temperatures, and the pressure drop for an existing heat exchanger; therefore the heat transfer surface area and the flow passage dimensions are available. The sizing problem, however, is concerned with the determination of the dimensions of the heat exchanger. In the sizing problem, an appropriate heat exchanger
Page | 24
type is selected and the size to meet the specified hot and cold fluid inlet and outlet temperatures, flow rates, and pressure drop requirements, is determined.
Figure 3.7 The Rating Program
There are some criteria for a successful heat exchanger design. These principles are given below. The process requirements which are accomplishing the thermal change on the streams within the allowable pressure drops, and retaining the capability to do this in the presence of fouling until the next scheduled maintenance period must be fulfilled. The exhaust gas heat exchanger must withstand the service conditions of the plant environment. The exchanger must be maintainable. In other words, a configuration that permits cleaning and replacement of any component that is especially vulnerable to corrosion, erosion, vibration, or aging, must be chosen. The heat exchanger should cost as little as possible provided that the above criteria are satisfied.
Limitations on the heat exchanger length, diameter, weight, and/or tube specifications due to site requirements, lifting and servicing capabilities must be all taken into consideration in the design.
There are some terms used in heat exchanger specification problems andtheir solutions, which are often confused. These are „rating‟, „design‟, and „selection‟. „Rating‟ is the computational process in which the inlet flow rates andtemperatures, the fluid properties, and the heat exchanger parameters are taken asinput and the outlet temperatures and thermal duty (if the exchanger length is specified) or the required length of the heat exchanger are Page | 25
calculated as output. In either case, the pressure drop of each stream will also be calculated. „Design‟ is the process of determining all essential constructional dimensions of an exchanger that must perform a given heat duty and respect limitations on shell-side and tube- side pressure drop. A number of other criteria are also specified, such as minimum or maximum flow velocities, ease of cleaning and maintenance, erosion, size and/or weight limitations, tube vibration, and thermal expansion. Each designproblem has a number of potential solutions, but only one will have the bestcombination of characteristics and cost. „Selection‟means choosing a heatexchanger from among a number of units already existing. Typically, these arestandard units listed
in catalogs of various manufacturers. Sufficient
manufacturer‟s data usually exist to allow one to select comfortably oversized exchanger with respect to both area and pressure drop.
Kern and Bell-Delaware methods will bepresented by performing Thermal Analysis and Hydraulic Analysis separately for the tube-side and for the shell-side. The rationale behind these choices is that the Kern method offers the simplest route.
3.5.1 Kern Method The first attempts to provide methods for calculating shell-side pressure drop and heat transfer coefficient were those in which correlations were developed based on experimental data for typical heat exchangers. One of these methods is the well-known Kern method, which was an attempt to correlate data for standard exchangers by a simple equation analogous to equations for flow in tubes. However, this method is restricted to a fixed baffle cut (25%) and cannot adequately account for baffle-to-shell and tube-to-baffle leakages.
Page | 26
3.6
Design Procedure of Heat Exchanger
Step 1
Step 10 Specification Define duty Make energy balance if needed To calculate unspecified flow Rates to temperatures
Decide baffle spacing and Estimate shell -side heat Transfer coefficient Step 11
Step 2
Calculate overall heat transfer Coefficient inculding fouling factors, U0,calc
Collect physical Properties
Step 3
no
O < (Uo,ass-U0,calc )/Uo,ass < 30 %
Assume value of overall Coefficient U0 ass Set Uo,ass = U0,calc
Step 4
yes
Step 12
Decide number of shell and Tube passes calculate ∆TIm Correction factor, F, and ∆Tm
Estimate tube and shell side Pressure drops
Step 5 Determine heat tranfser area Required A0 = q/Uo,ass ∆Tm
no
Pressure drops Within specification?
Step 6 Step 13
Decide type,tube,size,material Layout assignfluids to shell or tube side
yes
Estimate cost of exchanger
Step 7 Calculate number of tubes Step 14
Step 8
yes Calculate shell diammeter
Can design be Optimized to reduce cost ?
no
Step 9 Estimate tube-side heat Transfer coefficient
Accept design
Figure 3.8 Design Procedures for Exhaust Gas Shell and Tube Heat Exchanger Page | 27
3.7
Problem Statement The basic problem is to design a shell and tube heat exchanger which uses the
automobile diesel engine exhaust to heat bio-fuel from known room temperature to the elevated temperature to make the fuel atomisation finer. Mf=13.50kg/hr T2=91 Deg c
T1=35 Deg c
Bio-diesel Meg =12.05kg/hr Exhaust gases
T1=327 Deg c
T2
Figure 3.9 Problem Statement Basic problem Design a shell and tube heat exchanger in which exhaust gases at the inlet of exchanger having a mass flow rate of 12.05 kg/h & shell side flow rate of Bio-diesel is 13.50 kg/hr. Exhaust gases composition are as follows
Table 3.1 Composition of exhaust gas in mole fraction GAS
COMPOSITION IN MOLE FRACTION
CO2
0.0836
N2
0.8235
CO
0.00027
O2
0.0926
Exhaust gases composition has given in table. Page | 28
It is assumed that the temperature of exhaust gas at the inlet is 327°C and the inlet temperature of bio-diesel is 35°C and outlet temperature is to be achieved 91°C.Exhaust gases are put in tubes whereas bio fuel in shell & counter flow has been chosen. The materials for the tube has taken stainless steel .which have thermal conductivity of 45 W/m°C(shown in table) Fouling factor has taken = 0.00035(from table7.5) Specific dimensions for this heat exchanger (from table) Inner diameter of tube = 20 mm Outer diameter of tube = 16.8 mm Length of tubes = 600mm Desired pressure should not get exceed from 10 psi for both shell and tube.
3.8
Specific Heat of Bio-Diesel (Fluid) Formula used From perry design hand book 6.0 addition CP(Biodiesel) = A/(0.80)15/2 + B(t-15) A = 0.450 B = 0.0007 t is the temperature of biodiesel CP(Biodiesel) = 2.413 KJ/KgK
3.9
Properties of Exhaust Gases Specific Heat of Exhaust Gas Formula used Cpig/R = A+BT+CT+DT-2 Overall specific heat of exhaust gases Cp = Cpig (CO2)+Cpig (N2)+Cpig (CO)+Cpig(O2) = 0.0909+0.8913+2.96×10-4+0.0933 Cp = 1.075 KJ/KgK Page | 29
Calculation of Viscosity (µ) For Exhaust Gases By Sutherland‟s formula µ =µo× (a/b) × (T/To) 3/2 a = 0.555 To+ C b = 0.555 T + C Here µ = viscosity in centipoise at input temperature µo= Reference viscosity at reference temperature T = Input temperature in Rankin To = Reference temperature in Rankin C = Sutherland‟s constant a & b = constant Overall viscosity of the exhaust gas (µ) µ = (0.02737+0.02908+0.02924+0.03424) ×10-3 µ = 0.02998×10-3 Ns/m2 For The Thermal Conductivity K Euckens equation ,Eucken (1911)
K = µ (Cpeg+10.4/M) Where µ = viscosity of the exhaust gas Cpeg= specific heat of exhaust gas M = molecular mass of the gas Overall thermal conductivity of gas K =ΣXiKi = 0.1190×0.00027 + 0.11685×0.8235 + 0.1650×0.0836 + 0.1438×0.0926 = 0.1233 Page | 30
3.10 Results Obtained Mass flow rate of exhaust gases (meg)
= 12.05 kg/hr
Mass flow rate of fluid (mf) = 13.50 kg/hr Exhaust gas inlet temperature (T1)= 327°c =600 K Bio-diesel inlet temperature (t1) = 35°c= 308K Bio-diesel outlet temperature (t2)= 91°c=364K Cp of exhaust gas in KJ/KgK= 1.075 KJ/KgK Cp of Biodiesel= 2.413 KJ/KgK Viscosity of gases (µ)= 0.02998×103 kg/ms Viscosity of Biodiesel (µ)= 0.0004 kg/ms Thermal conductivity of gases (Keg)= 0.1233 W/mC Thermal conductivity of Biodiesel(Kf)=0.154 W/mC = 2.496 kg/m3
Density of the gas (eeg) Density the biodiesel (ef)= 736.8 kg/m3 Now Applying the Heat Balance Equation
Heat given by exhaust gas = Heat taken or receive by fluid meg×Cp×∆T = mf×∆t×Cp 12.05 Х 1.075(327 – T2) = 13.50 Х 2.413(91 – 35) So from here, T2 = 175.60 C = 448.60 K is outlet temperature of the cold fluid
Duty Q = meg × Cp×∆T (Q)Duty = 12.05/3600 ×1.075× (327 - t2) Q =0.544 KW Page | 31
Overall Heat Transfer Coefficient Heat U = 15 w/m2C (assumed) Exchanger Type & Dimensions The general equation for heat transfer across a surface is: Q = U×A×∆T Q = Heat transferred per unit time, W U = The overall heat transfer coefficient, W/m²°c A = Heat transfer area, m² Now logarithmic mean temperature difference(θm) θm = θ1- θ2/ ln(θ1 - θ2) θ1= 236deg C θ2= 140.6 degC θm = 184.20deg C θm = log mean temperature difference, T1 = hot fluid temperature, inlet, T2 = hot fluid temperature, outlet, t1 = cold fluid temperature, inlet, t2 = cold fluid temperature, outlet.
Temperature Ratios R= (T1-T2)/(t2-t1) =2.70 S =(t2-t1)/ (T1-T2)=0.1917 R is equal to the shell-side fluid flow-rate times the fluid mean specific heat; divided By the tube-side fluid flow-rate times the tube-side fluid specific heat. S is a measure of the temperature efficiency of the exchanger. From Figure 6.1 of COULSON & RICHARDSONS Page | 32
The correction factor is given by Ft= 0.96 θm =0.96×184.20 θm=176.83degC But, Q = U × A × θm So, A = 0.2066 m² For tubes, (D0) external diameter of the tube = 20 mm [From Table 6.4] (Di) internal diameter of the tube = 16.8 mm [From Table 6.4] Length of heat exchanger (L) = 600 mm Thickness of tube (t) = 1.6 mm Area of one tube, At = π ×Do × L = 0.03768 m² So number of tubes, Nt = .2066/0.03768 =6 Single pass with triangular pitch can be preferred:
Tube cross sectional area = π/4 (16.8 Х 10⁻³)² = 2.21 × 10⁻⁴ m² Page | 33
Tube side, Volumetric flow rate = 12.05/ (3600 × 2.496) = 1.341 × 10⁻³ m3/s Area per pass = no. of tubes in a single pass × tube cross sectional area = 6×2.21×10-4 = 1.326 ×10-3 m2 Tube Side Velocity (ut) = 1.341 × 10-3/ (1.326×10-3) (ut ) = 1.01 m/s The velocity is satisfactory, between1to 2 m/s, but may be a little low. This will show up when the pressure drop is calculated.
Bundle and Shell Diameter For single pass Table 6.3 from COULSON& RICHARDSONS K1=0.319 Triangular pitch = 1.25D0 n1= 2.142 Bundle diameter, Db =D0× (no. of tubes/K1)1/n1 Db = 20 ×10-3(6/0.319) (1/2.142) = 0.07868 m Db = 78.68 mm Shell Side Diameter Ds = Db+ Clearance Clearance =10% of Db Ds = 78.68+7.868 Ds = 86.54 mm
Page | 34
Pitch Pt= 1.25D0 Pt = 25 mm
Tube side heat transfer coefficient, Reynolds number, Re = ρ × V × Di/µ Re = 2.496×1.01×16.8×10-3/0.02998×10-3 Re= 1412.6 Prandtl number, Pr = µ×cp/Keg Pr = 0.2613 Length to diameter ratio, L/Di = 600/16.8 = 35.71 Heat transfer factor, From the Figure6.2 Jh = 2.01 × 10⁻2 Nusselt number, Nu = Jh×Re×(Pr)1/3 Nu = 2.01×10-2×1412.6×(0.2613)1/3 Nu= 18.15 (Heat transfer Coefficient for tubes), hi = Nu× (K/Di) hi = 18.15×(0.1233/16.8×10-3) = 133.20 W/m²C
Shell side heat transfer coefficient, Bundle diameter Db = 78.68 mm Shell diameter Ds = 86.54 mm
Page | 35
Baffle spacing = Ds/5 =86.54/5 = 17.30mm
Shell side cross flow area, As = (pitch – D0)/pitch ×Ds ×baffle spacing = 299.42mm2 = 2.99×10-4m2 Equivalent diameter, De = 1.10/16 × (Pt²- .917Do²) De = 17.25 mm Volumetric flow rate, Bio diesel mass flow rate = 13.50 Kg/hr Volumetric flow rate = 13.50 ×1/(3600 ×736.8) = 5.08 × 10-6 m³/s
Shell side velocity = volumetric flow rate/As = 0.0169m/s Reynolds number, Re = ρ × V × De/µ Re = (736.8 × 0.0169 × 17.75 × 10-3)/.00045 = 491.15 Prandtl number, Pr = µ ×Cp/Kf Pr = (0.00045 × 2.403× 10³)/0.154 Pr = 7.050 Page | 36
Heat transfer factor, Use segmental baffles with a 25% cut. This should give a reasonable heat transfer coefficient without too large a pressure drop.From figure 7.3 Jh = 4.4 Х 10-2 Heat transfer coefficient, hs =(K/De) ×Jn×Re×Pr1/3 hs = (0.154×103/17.75) ×4.4 Х 10-2×491.15 × (7.050)1/3 hs = 359.49 W/m²C Overall Heat Transfer Coefficient 1/Uo = (1/hi + fouling factor) × (D0×Di) + {D0×ln(D0/Di)}/2×thermal conductivity for the tube + 1/hs 1/U0=(1/133.20+0.00035)×(20/16.8)+{20ln(20/16.8)}/2×45 +1/359.49 U0 = 20.00 w/m2 C This is above the initial estimate of 15 w/ m2 C,this will prove to be a good heat exchanger from the economic point of view and might fullfill the basic need of designing an efficient heat exchanger.
Now Check the Pressure Drop No. of tube = 6 Pass =1 Inner diameter of tube (Di) = 20 mm Outer diameter of tube (D0) = 16.8 mm ut= 1.01 m/s Re =1.412 ×103 Pr = 0.2613 Jf (from Figure 7.4) =5.4×10-2
Page | 37
Tube Side Pressure Drop ∆Pt= no. of pass [8×Jf×(L/Di)×(µ/µw)-m + 2.5](e×Ut2/2) =1×[8×5.4×10-2×(35.71)+2.5]×(2.496×1.012/2) ∆Pt= 22.82 N/m2 Shell Side Pressure Drop ∆Ps= 8×4.4×10-2× (86.54/17.75)×(600/16.8)×(736.8×.01462/2) ∆Ps = 4.18 N/m2 The pressure drop is within limits, so the above dimensions might have taken.
Page | 38
CHAPTER 4 RESULT & DISCUSSION
4.1
Cost Estimation of Heat Exchanger
Raw material: 1) S.S pipe Ф 100×1×650 mm. qty – 1
1760 Rs
Ф 20 ×1.5 ×3900 mm. qty – 1
3276 Rs
2) M.s plate 150×150 × 15 qty – 2
220 Rs
3) Nut and bolt M6 ×1.5 inch – (1kg)
100 Rs
4) Nozzle ½ inch qty – 2
50 Rs
5) Baffle plate
200 Rs
6) Other parts
500 Rs
Total amount
6316 Rs
Machining cost: 1) Lathe at Rs 100/hr
200 Rs
2) Drill machine at Rs 80/hr
160 Rs
3) Brazing
250 Rs
4) Assemble cost
150 Rs
5) Over heads (10% of total cost)
707.60 Rs
6) Cost of heat exchanger
7783.60 Rs
NOTE: -
Cost is taken on the basis of market survey
-
Cost of product and machining may vary
-
Cost of heat exchanger is not actual Page | 39
4.2
Payback Period The payback period is how long it will take to recover money invested in a project,
and the so-called straight-payback-period calculation is the simplest way of determining the project‟s investment potential. 1) Total cost of heat exchanger
7800 Rs(approx.)
2) Additional running cost
1000 Rs/year
3) Energy consumption without shell and tube : Electrical energy consumption
0 .750 kwh/hr
No. of operating shift
250
No. of hr/shift
18
Cost of electrical
3.08 Rs/hr
Annual cost of energy without heat exchanger
12825Rs
4) energy consumption with shell and tube : Electrical energy consumption
0 .544Kwh/hr
No. of operating shift
250
No. of hour/shift
18
Cost of electricity
3.80 Rs/kwh
Annual energy cost with heat exchanger
9302.4 Rs
5) Saving / year = 3 – (4+2) = 12825 – 9302.4 – 1000 = 2522.6 6) Simple payback period = 7800/2522.6 = 3.09
Page | 40
The result of an economic evaluation of the energy saving using shell and tube heat exchanger as shown above The actual material and labour cost (cost including all variations on the bases of market survey) incurred to retrofit the shell and tube heat exchanger are reflected in table .which account to 7800 Rs and with the energy saving minimum additional running costs of 2500 Rs (watch) yielded a simple payback period of 3.09 year that is the period the heat exchanger will take recovery of money invested in a project. So the result shows that the heat exchanger installation would be a good idea at the outlet of exhaust pipe.
Page | 41
Figure 4.1 Wireframe View of Exhaust Gas Heat Exchanger
Figure 4.2 Solid View of Exhaust Gas Heat Exchanger
Page | 42
CHAPTER 5 CONCLUSION
From the above it is found that the exhaust gases at a very high temperature can be utilised to reduce the viscosity of biodiesel through heat exchanger, the exhaust gases at a low temperature at the outlet of the heat exchanger can also help in reducing the health hazards causes due to the direct exposure of high temperature exhaust gases in atmosphere. The use of shell and tube type heat exchanger for this application is also economic than any other type of heat exchanger as mentioned in this report. The design of such type of heat exchanger is compact and might be facilitate at the exhaust of an diesel engine. The problem might be its length and its compactness from the point of view of manufacturing. But if all these encountered then the above design can be useful and efficient for above purpose.
Page | 43
Mass flow rate of exhaust gases (meg)
12.05 kg/hr
Mass flow rate of fluid (mf)
13.50 kg/hr
Exhaust gas inlet temperature (T1
600 K
Bio-diesel inlet temperature (t1)
308K
Bio-diesel outlet temperature (t2)
364K
Cp of exhaust gas in KJ/KgK
1.075 KJ/KgK
Cp of Biodiesel
2.413 KJ/KgK 0.02998×103 kg/ms
Viscosity of gases (µ) Viscosity of Biodiesel (µ)
0.0004 kg/ms
Thermal conductivity of gases (Keg)
0.1233 W/mC
Thermal conductivity of Biodiesel
(Kf)
0.154 W/mC
Density of the gas (eeg)
2.496 kg/m3
Density the biodiesel (ef)
736.8 kg/m3
Duty LMTD Heat transfer area A (D0) External diameter of the tube
0.544 KW 184.20 C 0.2066 m2 20 mm
(Di) Internal diameter of the tube
16.8 mm
Length of heat exchanger
600 mm
Number of tube
6 Page | 44
Tube side velocity ut Shell diameter Ds Pitch Pt Heat transfer Coefficient for tubes, hi Equivalent diameter De
1.01 m/s 86.54mm 25 mm 133.20 W/m²C 17.25 mm
Overall Heat Transfer Coefficient Uo
20.00 w/m2
Heat transfer coefficient at shell side
359.49 W/m²C
Tube Side Pressure Drop ∆Pt
22.82 N/m2
Shell Side Pressure Drop ∆Ps
4.18 N/m2
Page | 45
CHAPTER 6 APPENDICES
PROPERTIES OF EXHAUST GASES SPECIFIC HEAT OF EXHAUST GAS Formula used All data of A, B, C and D taken from table 7.6 Cpig/R = A+BT+CT+DT-2 For CO2 A = 5.457 B = 1.045×10-3 C = 00 D = -1.157×105 Cpig/R = 5.7627 Mole fraction of CO2 = 0.0836 Value of R = 0.1889 KJ/KgK Cp=0.0909 KJ/KgK For N2 A = 3.280 B = 0.593×10-3 C = 00 D = 0.040×105 Cpig/R = 3.6469 Mole fraction of N2 = 0.0836 Value of R = 0.2968 KJ/KgK Cpig = 0.8913 KJ/KgK For CO A = 3.376 B = 0.557×10-3 Page | 46
C = 00 D = -0.031×105 Cpig/R = 3.70159 Mole fraction of CO = 0.00027 Value of R = 0.2968 KJ/KgK Cpig = 2.96×10-4 KJ/KgK For O2 A = 3.639 B = 0.506×10-3 C = 00 D = -0.277×105 Cpig/R = 3.8795 Mole fraction of O2 = 0.0926 Value of R = 0.2598 KJ/KgK Cpig = 0.0933 KJ/KgK Overall specific heat of exhaust gases Cp = Cpig (CO2)+Cpig (N2)+Cpig (CO)+Cpig(O2) = 0.0909+0.8913+2.96×10-4+0.0933 Cp = 1.075 KJ/KgK CALCULATION OF VISCOSITY (µ) FOR EXHAUST GASES By Sutherland‟s formula µ =µ0× (a/b) × (T/T0) 3/2 a = 0.555 T0+ C b = 0.555 T + C Here µ = viscosity in centipoise at input temperature µ0= reference viscosity at reference temperature T = input temperature in Rankin T0 = reference temperature in Rankin C = Sutherland‟s constant a & b = constant
Page | 47
For CO2 Temperature = 527.67 R µ0= 0.01480 centipoise C = 240 a = 530.21 b = 839.54 From Sutherland‟s formula µ (CO2) = 0.02737 centipoise 1 centipoise = 10-2 poise µ= 0.02737×10-3 Ns/m2
For N2 Temperature = 527.67 a = 411.24 b = 710.56 C = 111 µ = 0.02908×10-3 Ns/m2
For CO Temperature = 527.67 a = 405.86 b = 717.54 C = 118 µ =0.03424×10-3 Ns/m2
For O2 Temperature = 527.67 µ0= 0.02018 centipoise C = 127 a = 526.05 b = 1080.27 From Sutherland‟s formula µ (O2) = 0.03424 centipoise Page | 48
µ (O2) = 0.03424×10-3 Ns/m2 Overall viscosity of the exhaust gas (µ) µ = (0.02737+0.02908+0.02924+0.03424) ×10-3 µ = 0.02998×10-3 Ns/m2 The following table gives the values of Sutherland's constant and the reference temperature and viscosity for the gases used in the LMNO Engineering calculation. Values of Sutherland's constant are from Crane (1988, p.A-5). The reference temperatures and viscosities were selected from CRC (1984, pp.F-42-44).
Table 6.1 Values of Sutherland‟s constant and reference temperature and viscosity Sutherland's constant,C
To (oR)
µo (centipoise)
standard air
120
524.07
0.01827
ammonia, NH3
370
527.67
0.00982
carbon dioxide, CO2
240
527.67
0.01480
carbon monoxide, CO 118
518.67
0.01720
hydrogen, H2
72
528.93
0.00876
nitrogen, N2
111
540.99
0.01781
oxygen, O2
127
526.05
0.02018
sulfur dioxide, SO2
416
528.57
0.01254
For The Thermal Conductivity (K) Euckens equation ,Eucken (1911) K = µ (Cp+10.4/M) For CO K (CO) = 0.1190
For CO2 K (CO2) = 0.02751(5.7627+10.4/44) Page | 49
For O2 K (O2) = 0.3424× (3.875+10.4/32) K (O2) = 0.1438 For N2 K (N2) = 0.2751× (5.7627+10.4/44) K (N2) = 0.11685 Overall thermal conductivity of gas K =ΣXiKi = 0.1190×0.00027 + 0.11685×0.8235 + 0.1650×0.0836 + 0.1438×0.0926 = 0.1233
Figure 6.1 Temperature correction factor: one shell pass; two or more even tube passes
Page | 50
Table 6.2 Conductivity of metals
Figure 6.2 Tube-side heat-transfer factor
Page | 51
Figure 6.3 Shell-side heat-transfer factors,segmental baffles
Figure 6.4 Tube-side friction factors
Page | 52
Table 6.3 Constants for use in equation
Table 6.4Standard dimensions for steel tubes
Page | 53
Table 6.5Foulingfactors (coefficients), typicalvalues
Page | 54
Table 6.6 Heat Capacities of Gases in the Ideal Gas State
. Page | 55
CHAPTER 7 BIBLIOGRAPHY
1. Ahmet Necati Ozsezen, Ali Turkcan and Mustafa Canakci,“Combustion analysis of preheated
crude
sunflower”,
Alternative
Fuels
R&D
Center,
Kocaeli
University,41040 Izmit,Turkey,10 November 2008. 2. Feng Yang;Xingan Yuan,Guiping Lin,May 2002,accepted in 11th October 2002 ,waste heat recovery using heat pipe heat exchanger for heating automobile using exhaust gas. www.sciencedirect.com 3. Seokhwan lee, Choongsik Bae,available 10th May 2007;Design of a Heat Exchanger to reduce the exhaust temperature in a S.I engine,www.sciencedirect.com 4. K P M Wipplinger, T M Harms, A B Taylor, Stainless steel finned tube heat exchanger design for waste heat recovery, Journal of energy in southern Africa Vol 17 No 2.May 2006. 5. Chemical engineering design, 4th edition, volume 6, Coulson & Richardson`s chemical engineering series; R.K.SINNOTT. 6. J.M smith, chemical engineering thermodynamic ,6th edition, Tata Mcgrawhill. 7. Ramesh K. Shah and Dusan P. Sekulic, “Fundamentals of heat exchanger of design” 203 Jon wiley & Sons.
Page | 56
