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1.1 Heat exchangers

A heat exchanger is an equipment which transfers energy from a hot fluid to cold fluid by virtue of temperature difference. The fluid may be separated by a wall to prevent mixing or they may be in direct contact. In heat exchanger, the temperature of each fluid varies as it passes through the exchanger and hence the temperature of the dividing wall between the fluids also changes along the length of the exchanger. The ability of heat exchanger to transfer heat from hot fluid to the cold fluid governs the thermal performance of the system.

1.2 TYPES OF HEAT EXCHANGERS

1.2.1 ACCORDING TO NATURE OF HEAT EXCHANGE PROCESS

i) Direct contact              ii) In-direct contact

i) Direct contact heat exchangers : In a direct contact heat exchanger the heat exchange takes place by direct mixing of hot and cold fluid and mass transfer simultaneously. Some common examples  are cooling tower , jet condenser etc.

ii) Indirect contact heat exchanger : In the indirect contact heat exchanger, the fluid streams remains separate and heat transfer continuously through an impervious dividing wall. Some common examples are Regenerators and Recuperators.

1.2.2 ACCORDING TO RELATIVE FLUID MOTION

Heat exchanger according to the flow direction are divided into three categories

i) Parallel flow    ii) Counter flow     iii) Cross flow

i) Parallel flow : When the streams of the two fluid (hot and cold) flow in same direction and follows the same path from entry to exit. Some common examples are oil coolers, oil heaters.

Fig. 1.1 Parallel flow heat exchanger

ii) Counter flow : When both the fluid stream travels in opposite direction. The hot and cold fluid enter at the opposite end.

                            Fig.1.2 Counter flow heat  exchanger

iii) Cross Flow : When one fluid flows perpendicular to the second fluid ; that is, one fluid passes through the tubes and second fluid passes around the tubes at  90''  angle.

.  

Fig.1.3 Cross flow heat exchanger

1.2.3 ON THE BASIS OF DESIGN AND CONSTRUCTIONAL FEATURES

i) Shell and tube type : It consists of bundle of tubes mounted in a cylindrical shell with the tube axis parallel to that of shell. One fluid passes through the bundles of tubes while the other passes through the shell or outside the tube. This type of heat exchanger are used where reliability and heat transfer effectiveness is important.

                              

                                                    Fig. 1.4 Shell and tube type heat exchanger

ii) Concentric tube heat exchanger : This type of heat exchanger has two concentric tubes carrying different fluids. The fluid direction can be same or opposite. The higher effectiveness can be achieved by cross flow or by using swirling flow.

Fig. 1.5 Concentric tube heat exchanger

iii) Compact heat exchanger : These type of heat exchangers have a very large surface area as compared to volume of heat exchanger. They are generally used where convective heat transfer coefficient of one fluid is smaller than the other fluid. Some common examples are plate fin and flattened fin heat exchanger .

Fig.1.5 Plate fin heat exchanger

1.3 PROCESS OF HEAT TRANSFER IN PLANE TUBE HEAT EXCHANGER

Heat transfer process in a plain tube heat exchanger occurs due to conduction process which takes place on the tube surface and the convection process which takes place between heated surface of plain tube and the fluid which is flowing inside it.

Heat transfer takes place due to breaking of hydrodynamic and thermal boundary layer on the inner side of the tube. The boundary layers offers resistance in the path of heat transfer between the fluid flowing through the tube surface and inner surface of the tube. Among various techniques which have been investigated for increment of heat transfer rates inside circular tubes, a wide range of inserts have been utilized. The various types of inserts include twisted tapes, delta winglet tape, perforated tapes, conical rings etc. These techniques give rise to prolonged residence time of the flow in the tube, thinning the boundary layer which leads to the increase in heat transfer inside the tube.

1.4 ADVANTAGES OF TUBE TYPE HEAT EXCHANGER

1) Less expensive as compared to plate type heat exchanger because the titanium plates which are used in plate type heat exchangers are costly.

2) Can be used with system at high operating temperature and pressure

3) Tubes leaks are easily located because the pressure test in tube type heat exchanger is easy as compared to plate type heat exchanger.

4) Pressure drop across a tube cooler is less because over tightening of clamping bolt results in high pressure drop in plate type heat exchanger.

1.5 DISADVANTAGES OF TUBE TYPE HEAT EXCHANGER

1) Heat transfer efficiency is less as compared to plate type

2) Clearing and maintenance is difficult since tube cooler requires enough clearance at one end to remove the corrosion inside it.

3) Capacity of tube cooler can not be increased.

4) Requires more space as compared to plate coolers because parts of tube cooler are large.

CHAPTER 2

LITERATURE REVIEW

Demand of energy is increasing day by day due to which conventional sources of energy are depleting at higher rates. Researches are going on to improve the performance of heat exchangers with the objective to reduce size as well as cost of heat exchangers. Several heat transfer augmentation techniques have been developed and applied to heat exchangers to date. One of the main techniques is known as passive technique in which heat transfer rate can be increased by using modified geometries or modified surfaces such as treated surface, rough surfaces, extended surfaces, coiled tube etc. Inserts are one of the main passive heat augmentation devices and can be of any geometries like twisted tapes, delta winglet twisted tapes, conical rings, v-nozzles, swirling jets etc. Inserts are used to enhance heat transfer rate with less rise in frictional loses. Kumar and Prasad [1] conducted the experiment on water to evaluate heat transfer and friction factor by using typical twisted tapes in solar water heaters having twist ratios (y) between 3 ' 12. The Reynolds number was varied in the range of 4,000 to 12,000 and it was found that as the twist ratio decreases, the Nusselt number and friction factor increases. Nusselt number was increased from 18 to 70% and pressure drop was increased from 87 to 132% by using simple twisted tape, while increment in thermal efficiency was 30%. Twisted tapes create turbulence and swirl flow inside the solar heaters due to which heat transfer enhancement increases.

 Fig.2.1 Twisted tape and tube geometry ( Kumar [1] )

Sarada et al. [2] performed the experiment for enhancement of heat transfer using various widths twisted tape inserts in a horizontal tube with air as a working medium. Widths of twisted tapes varied from 10 mm to 22 mm and found that as the width of twisted tape increases the heat transfer rate also increases. For width of 10 mm, increment in Nusselt number was found to be 2 to 8%, for 14 mm increment in Nusselt number was 6 -12%, and for 18 mm increment in Nusselt number was 9 -19%, for 22 mm increment in Nusselt number remained 14 '27%. The Reynolds number was varied from 6,000 to 13,500. Twisted tapes leads to enhanced heat transfer rate by generating swirl or secondary flows increasing the flow velocity due to the tube partitioning and blockage. .

Fig. 2.2 Various widths twisted tape ( Sarada et al. [2] )

Noothong et al. [3] had performed experiments to investigate the effect of twisted tape inserts on heat transfer in a tube using water as a testing fluid. The Reynolds number was varied from 2,000 to 12,000. Increment in Nusselt number at (y) of  5 was 188% times of plain tube, at  (y) of 7 Nusselt number was 159% times of plain tube. The twisted tape causes swirl and pressure gradient in radial direction due to which the boundary layer along the tube wall becomes thinner resulting in more heat flow through the fluid.  As the twist ratio decreases Nusselt number and friction factor increases due to swirl flow created by twisted tapes which increases residence time of fluid .

Fig. 2.3 Inner tube fitted with twisted tape  with pitch ratio (y) of 5 and 7 ( Noothong et al. [3] )

Eiamsa-ard et al. [4] performed experimental investigation of heat transfer and friction flow in a circular tube using regularly spaced twisted tape with water as a working medium. They uses full length twisted tape and twisted tape with various space ratio. For full length twisted tape increment in Nusselt number at pitch ratio (y) of 6 was 179 % times of plain tube, at (y) of 8 Nusselt number increased by 143% times of plain tube. The variation of Nusselt number in regularly spaced twisted tape with space ratio (s) of 0 - 3 and at twist ratio (y)  of 6 was analyzed. At space ratio of (s) equals to 0 the Nusselt number was increased to 179 % of plain tube, at (s) equals to 1 the Nusselt number was increased to 160 % of plain tube, at (s) equals to 2 the Nusselt number was increased to 142 % times of plain tube, at (s) of  3 the Nusselt number was increased to 121% times of plain tube. There was a decrement in friction factor with increasing space ratio, for space ratio (s) of 1 decrement in (f) was 15 %, for space ratio (s) of 2 decrement in (f) was 39 %, for space ratio (s) of 3 decrement in (f) was 88%. The result reveals that Nusselt number and friction factor increases by decreasing space ratio because as we decrease the space ratio more turbulence is created by twisted tape.  

Fig. 2.4 Regularly spaced twisted tape ( Eiamsa-ard et al. [4] )

    Eiamsa-ard and Promvonge [5] investigated the effects of alternate clockwise and counter clockwise twisted tape in their experimental investigation using water as a working fluid having twist ratios (y) of  3 , 4 and 5 and twist angles ('') of 30 '' , 60 ''  and 90 ''  . There is an increment in Nusselt number from 12.8 ' 41.9 % that of typical twisted tape, while Reynolds number was varied from 3,000 to 27,000. The friction factor (f) increases from 19.6 ' 32.8 % that of typical twisted tape and thermal performance factor  ('')  increases from 1.19 ' 1.28 . The result reveals that heat transfer rate of alternate clockwise and counter clockwise twisted tape increases with decrease of twist ratio and increase of twist angle. This modified twisted tape is designed to offer periodic swirl direction along the test tube which is expected to provide better mixing of fluid.

Fig. 2.5 Alternate clockwise and counter-clockwise twisted tape ( Eiamsa-ard et al. [5] )

Fuskele and Sarviya [6] performed experimental investigation of heat transfer and friction factor using twisted dense wire mesh having twist ratio (y) of 5 and 7. At twist ratio (y) of 5, Nusselt number was increased to 1.69 times of plain tube, at twist ratio (y) of 7 Nusselt number was increased to 2.09 times of plain tube. Water was used as a working medium and Reynolds number was in the range of 4,000 to 30,500. At twist ratio (y) of 5 friction factor increases to 4.3 times of plain tube, at twist ratio (y) of 7 friction factor increases to 4 times of plain tube. At twist ratio (y) of 5 thermo-hydraulic performance factor increases from 1.23 ' 1.53 times of plain tube , at twist ratio (y) of  7 thermo-hydraulic performance factor increases from 1.16 to 1.49 times of plain tube. The result obtained lead to the conclusion that higher heat transfer rates can be achieved by using porous inserts at the expense of reasonable pressure drop.

 Fig. 2.6 Twisted dense wire mesh with twist ratio(  y ) of 5 and 7 ( Fuskele et al. [6] )

    Eiamsa-ard et al. [7] performed the experiment in which they investigated the effect of helical screw tape with and without core rod inserts in double pipe heat exchanger device on heat transfer and flow friction characteristics.Water was  used as a working medium and the Reynolds number was varied in the range of 2,000 to 12,000. The stainless steel helical screw tape has the geometrical configuration of width 17 mm and clearance of tube surface was 4 mm. The loose fit helical screw tape with and without core rod were used as an inserts. It is found out that increment in Nusselt number by using loose fit helical screw tape with and without rod was 230% and 340% that of plain tube. Friction factor (f) was 50% less by using helical screw tape without core rod than that for one with core rod. While the enhancement efficiency varies between 1 and 1.17, 1.98 and 2.14 for the tape with and without rod.

Fig. 2.7 a) Helical screw tape with-out rod ( Eiamsa-ard et al. [7] )

Fig. 2.7 b) Helical screw tape with rod ( Eiamsa-ard et al. [7] )

  Eiamsa-ard et al. [8] investigated the effect of delta winglet twisted tape on heat transfer in a circular tube having twist ratio (y/w) of 3, 4 and 5 and depth ratio (d/w) of 0.11,  0.21 and 0.32 . An oblique delta winglet twisted tape and straight delta winglet twisted tape were used as an insert. Water was used as a working fluid and the Reynolds number was varied between 3,000 to 27,000. The result revealed that there is an increment in heat transfer coefficient and friction factor by reducing the value of twisted tape and increasing depth of cut. Oblique delta winglet twisted tape is better turbulator than straight delta winglet tape. By using oblique delta winglet there is an increment in Nusselt number from 1.04 ' 1.64 times that of typical twisted tape  and friction factor increases from  1.09 ' 1.95 times that of typical twisted tape while enhancement efficiency increases ( '') from 1.05  to 1.13 times that of typical twisted tape.

Fig. 2.8 a) typical twisted tape b) straight delta winglet twisted tape c) oblique delta winglet twisted tape ( Eiamsa-ard et al. [8] )

  Eiamsa-ard et al. [9] investigated the effect of peripherally cut twisted tape insert on heat transfer and thermal performance factor in laminar and turbulent tube flows having constant twist ratio (y/w) of 3 , depth ratio (d/w) of 0.11, 0.22,  0.33 and width ratio (w/W) of 0.11, 0.22 and 0.33. Tests were performed with Reynolds number in a range from 1,000 to 20,000 and water was used as a working fluid. The results show that as the depth ratio increases and width ratio decreases the heat transfer enhancement was increased . There is an increment in Nusselt number upto 2.6 times in turbulent regimes and 12.8 times in laminar regimes that of plain tube. While thermal performance factor ('') increases to 1.29 in turbulent regimes and 4.88 in laminar regimes. Increment in heat transfer is due to higher turbulence intensity of fluid in the surrounding area of tube wall created by peripherally cut twisted tape.

Fig. 2.9 Test tube with peripherally cut twisted tape ( Eiamsa-ard et al. [9] )

 Eiamsa-ard et al. [10] conducted experiment on thermohydraulic investigation of turbulent flow through a round tube equipped with twisted tape consisting of centre wings and alternate axes having a constant twist ratio (y/W) of 3, using water as a working fluid. The effects of other three types of  twisted tapes including twisted tape with wings alone (WT) , twisted tape with altenate axes alone (T-A) and typical twisted tape (TT) were also studied for comparision. The Reynolds number was varied between 5200 and 22,000 . The wings were generated along the centre line of the tape with three different angle of attack, ('') of 43'' ,53'' ,74'' . The heat transfer rate in a tube fitted with twisted tape with wings and alternate axes was higher than those in the tube fitted with twisted tape with wings, twisted tape with alternate axes and plain tube and the heat transfer rate increases with increasing angle of attack. Increment in mean value of Nusselt number provided by twisted tape with wing and alternate axes with ('') 74 '' is 17.7 % higher than in the tube with twisted tape with wings at same angle of attack , 20.8 % higher than those in the tube with twisted tape with alternate axes and 62% higher than those in the tube with twisted tape. Mean value of Friction factor (f) provided by twisted tape with wings and alternate axes at ('') of 74''  also increases to 30.6%  than in tube with twisted tape having wings with same angle of attack , 53% higher than those in the tube with twisted tape alternate axes and 123% higher than those in tube with twisted tape. Increment in mean thermal performance factor ('') provided by twisted tape with wings and alternate axes at ('') of 74''  is 7.8% higher than those in tube with twisted tape having wings at same angle of attack , 4.9% higher than those in tube with twisted tape alternate axes, and 24% higher than those in tube with twisted tape.

Fig. 2.10 Twisted tape with centre wing and alternate axes ( Eiamsa-ard et al. [10] )

  Bhuiya et al. [11] performed experiment on heat transfer and friction factor in turbulent flow through a tube fitted with perforated twisted tape having twist ratio of 1.92. Perforated twisted tape with four different radius of porosities (Rp) of 1.6%, 4.5%, 8.9% and 14.7% were used. Air was used as a working fluid and the Reynolds number was varied between 7200 to 49,800. The twisted tape with porosity of 4.5% provided the higher heat transfer rate than those of the other tapes of porosities 1.6%, 8.9% and 14.7%, this is because higher perforation causes less swirling effect. Increment in Nusselt number was 110% - 340% higher than those of plain tube. While the increment in friction factor (f) with perforated twisted tape inserts were 110-360% higher than those of the plain tube. The maximum thermal performance of 59% was achieved with the use of perforated twisted tape at porosity of 4.5% and at Reynolds number of 7525.

Fig. 2.11 Perforated twisted tape ( Bhuiya et al. [11] )

  Promvonge [12] investigated the heat transfer characteristics in a round tube with conical ring inserts. Conical rings with three different diameter ratios of the ring to tube diameter (d/D) of 0.5,  0.6, 0.7 are used in the experiment and for each diameter ratio, the rings are placed with three different arrangements converging conical rings, diverging conical rings and converging ' diverging array. Air was used as a working fluid and the Reynolds number was varied in the range of 6,000 to 26,000. Increment in Nusselt number for Converging conical rings array was upto 197 %, for Diverging conical rings array upto 333% and for Converging-Diverging conical rings array upto 237%. The reason behind the increment is the boundary layer disruption which causes a better chaotic mixing between the core and wall regions thus enhancing a convective process. The friction factor decreases at around 50% and 75 % for using (d/D) of 0.6 and 0.7 instead of (d/D) of 0.5. The enhancement efficiency ('') varies between 1.8 to 0.93 for Diverging conical rings array at (d/D) of 0.5 . For Converging conical rings array the efficiency ('')  varies between 1.12 ' 0.885 at (d/D) of 0.5 and for Converging Diverging rings array the efficiency ('') varies between 1.41 ' 0.89 at (d/D) of 0.5 . As the diameter ratio is increased there is a decrement in Nusselt number , friction factor and enhancement efficiency. The use of Diverging conical rings array provides better heat transfer than that of Converging conical rings or Converging Diverging conical rings arrays at similar d/D ratio this is due to higher re-circulation and higher contact surface area between the fluid and the heating wall surface when the fluid decelerates from Diverging conical rings array.

Fig. 2.12 a) Diverging conical ring arrangement b) converging conical ring arrangement c) converging-diverging conical ring arrangement ( Promvonge [12] )

Promvonge and Eiamsa-ard  [13] performed the experiment on heat transfer characteristics in a tube with combined conical ring and twisted tape. Two twisted tape of twist ratio (y) of 3.75 and 7.5 were used in the test set up. Air as a working fluid was used and the range of Reynolds number was varied between 6,000 to 26,000. At pitch ratio (y) of 3.75 increment in Nusselt number was 367% over the plain tube and at pitch (y) of 7.5 increment in Nusselt number was 350% over the plain tube. The average increase in friction factor is 145 times above the plain tube. The heat transfer and friction factor increased due to turbulence and swirl flow created near the wall region by using combination of twisted tape and conical rings. Enhancement efficiency of 1.96 is found for using conical ring and twisted tape at pitch ratio (y) of 3.75.

Fig. 2.13 Test tube fitted with conical ring and twisted tape ( Promvonge and Eiamsa-ard [13] )

Promvonge and Eiamsa-ard [14] investigated the heat transfer in a circular tube with free spacing snail entry and conical nozzle turbulators having pitch ratio of 2, 4 and 7 using air as a working medium . The Reynolds number range was varied from 8,000 to 18000. Increment in Nusselt number at (y) of 2 was 315% times of plain tube,  at (y) of 4 increment in Nusselt number was 300% times of plain tube and at (y) of 7 increment in Nusselt number was 285% times of plain tube. The heat transfer rate at lower pitch ratio is greater than that at the higher ones because the turbulence intensity and flow path obtained for lower pitch ratio is greater than higher ones. The increase in pressure loss for pitch ratio of 2, 4 and 7 are 87, 75 and 43 times of plain tube.The variations of enhancement efficiency for Reynolds number ranging from 5,000 to 18,000 are between 0.76 and 0.93 , 0.7 and 0.85 , 0.67 and 0.8 for pitch ratio of 2.0, 4.0 and 7.0.

Fig. 2.14 Test tube fitted with c-nozzle turbulator and snail with free space entry (Promvonge and Eiamsa-ard [14] )

  Promvonge and Eiamsa-ard [15] performed the heat transfer enhancement in a tube combined with conical ' nozzle and swirl generator using air as a working fluid . The Reynolds number was varied in the range of 8,000 to 18,000. Three different pitch ratio of conical nozzle arrangement was used in the set up with pitch ratio of 2, 4 and 7. It was found that the increase in heat transfer rate with reducing pitch ratio is due to the higher turbulence intensity . The Nusselt number inceased upto 300 times of plain tube at pitch ratio of 2, at pitch ratio of 4 the Nusselt number is increased to 257 times of plain tube and at pitch ratio of 7, the Nusselt number was increases to 236 times of plain tube. The increase in pressure loss is 71 times that of plain tube the reason behind pressure loss is the secondary flows occurring as a result of the interaction of the pressure forces with inertial forces in  the boundary layer.The enhancement efficiency varied between 0.75 and 0.89 , 0.73 and 0.86 , 0.70 and 0.83 for pitch ratio of 2, 4 and 7. The turbulators are applicable at low Reynolds number .

Fig. 2.15 a) conical nozzle arrangement b) conical nozzle with swirl generator ( Promvonge and Eiamsa-ard [15] )

  Kongkaitpaiboon et al. [16] performed experimental investigation of heat transfer and friction in a tube with perforated conical rings . The perforated conical rings are of three different pitch ratios (p/D) of  4, 6 and 12 and three different numbers of perforated holes (N) of 4, 6 and 8. Air was used as a working fluid and Reynolds number was varied in the range of 4,000 to 20,000. The mean heat transfer rates obtained from using the perforated conical rings with pitch ratio (P.R.) of 4, 6 and 12 are found to be 185%, 145% and 86% over the plain tube. This is due to interruption of flow by the turbulators which result in destruction of thermal boundary layer near the wall region.  The Nusselt numbers in the tube equipped with Perforated Conical Rings with number of holes (N) of 4, 6 and 8 holes are 90-239% , 69-220% and 65-172% higher than those in plain tube. While the friction factors (f) in the tube with Perforated Conical Rings are significantly lower than those in the tube with typical conical rings, which are around 72.2% for pitch ratio (PR) of 4, 68.1% for pitch ratio (PR) of 6 and 72.5% for pitch ratio (PR) of 12 as the pitch ratio increases the friction factor decreases. The friction factor of using the Perforated conical rings with (N) of 4, 6, and 8 are 57.2%, 73.6% and 82% lower than that of plain tube. The maximum thermal performance factors of tube fitted with Perforated conical rings at pitch ratio (PR) of 4, 6 and 12 are 0.92, 0.87 and 0.79. The performance factor of conical rings are lower than Perforated conical rings and this confirms the greater benefit of the use of Perforated conical rings as energy saving device than the conical rings.

Fig. 2.16 Perforated conical rings ( Kongkaitpaiboon et al. [16] )

The literature review related to heat transfer performance and pressure drop in heat exchanging devices shows that various number of geometries and tube inserts have been investigated till date. Heat transfer and fluid flow characteristics in terms of Nusselt number, friction factor over a range of Reynolds number is observed to determine the performance of heat exchanging device. An admirable heat exchanger must have high heat transfer rate and thermo-hydraulic performance at the expense of least pumping power.

With the enhancement in heat transfer rates by using different inserts in circular tubes, the frictional losses increases apparently. In order to reduce these frictional losses, recent study reveals that solid conical rings, peforated conical rings and slotted conical rings leads to higher heat transfer augmentation than twisted tape.

 The present work is taken up to investigate the effect of diverging conical rings with circular and slotted perforation on its circumference, on heat transfer and pressure drop characteristics of a flow through circular tube. The experimental data are to be collected for different types of diverging solid conical rings, perforated diverging conical rings and diverging slotted conical rings with different pitch ratio, number of holes and number of slots. The range of Reynolds number was varied from 4,000 to 11,000. The data for smooth tube is also collected to facilitate the comparison of results in case of perforated diverging conical rings. Pressure drop and heat transfer in terms of friction factor (f) and Nusselt number (Nu) are discussed with respect to Reynolds number and pitch ratio. The thermo-hydraulic performance factor is also observed and collected for different geometries of cones to determine the output or gain of heat exchanger.

The major objectives of present research work are as follows :-

1. To collect the experimental data pertaining to heat transfer and friction in a circular tube with different configurations of solid conical ring, perforated conical rings and slotted conical ring.

2. To discuss the heat transfer and pressure drop in terms of Nusselt number and friction factor with respect to Reynolds number.

3. To examine the thermo-hydraulic performance for different geometries of solid conical ring, perforated conical rings and slotted conical rings over the entire range of Reynolds number .

'

CHAPTER 3

EXPERIMENTAL TEST FACILITY

Survey of the literature shows that the boundary layer developed over the surface of the tube plays a vital role in the convection process of heat transfer in heat exchanger tubes. The boundary layer which is developed over the surface transforms into static boundary layer due to the presence of viscous forces near the surface of the tube. Turbulence which is created inside the tube due to the presence of inserts, diminishes the static boundary layer and helps in proper mixing of the fluid between the surface and the core region of the tube. To know the behaviour of heat transfer and friction characteristics of circular tube fitted with conical inserts, experimental approach is adopted in the present work. Data in the form of surface temperatures,  inlet and outlet temperatures of the fluid, flow rate and pressure difference of the tube are collected under different operating conditions.

3.1 CONSTRUCTIONAL DETAILS OF THE EXPERIMENTAL TEST SET UP

The experimental set up consists of  water tank of capacity 200 Litre which is placed at a height of 2.5 m  from the ground. The inlet section consists of Galvanized iron pipe of length (1500 mm). Before the inlet section, ball valve and rotameter are attached . The ball valve controls the flow of water into the test section and rotameter measures the volume flow rate of water. Test section consists of copper pipe of length 1,000 mm having inner diameter of 24 mm and outer diameter of 25 mm, thickness (t) 1.5 mm. In order to measure the temperature, six T- type thermo-couples are attached to the surface of copper pipe at a distance of 160 mm. Nichrome wire is wound around the tube to maintain uniform heat flux over the entire test section. Heating element receives power through variac transformer to regulate voltage and thereby heat flux is controlled. Finally the pipe is insulated with the glass wool insulation followed by insulating foam. The exit section consists of galvanized iron pipe.  Two thermocouples are placed inside the tube at inlet and outlet of the test section. Temperature indicator is used to display the temperatures at various locations in the test section . Micro-manometer is used to measure the pressure difference across the test section.

Fig. 3.1 Schematic Diagram of the experimental setup.

Fig 3.2 Photographic view of experimental setup

 3.1 COMPONENTS OF EXPERIMENTAL SET-UP

3.1.1  CONTROL PANEL

i) Ammeter :- It is used to measure the current in the circuit . It measures the current upto 5 amperes.

ii) Voltmeter :- It is an instrument that measures the difference in electric potential between two points in an electric circuit.

iii) Variac :- It is used to vary the voltage across the circuit. The range of variac  is from 0 to  260 volts.The variac is connected to the heating element that controls the surface heat flux of the test section.

 iv) Temperature Indicator :- It is used to display the temperature of water at the inlet, exit and tube surface with the accuracy of 0.1''  C.

Fig 3.3 Control panel with temperature indicator

3.1.2  Rotameter and Glove valve :- is an assembly which is used to regulate and measure the fluid flowrate in a closed tube. The rotameter consists of tube and a float. The float responses to the change in  flow rate. The flow rate is regulated with the help of a glove valve. At higher flow rate float is raised to a greater height.

Fig.3.4 Rotameter and glove valve

3.1.3 Micro-manometer :- It is used to measure the pressure head in millimeter of Hg which is used to determine the friction factor. Digital micro-meter is used in the set up to measure the pressure difference between inlet and outlet of the test section. It is capable to measure small difference in pressure head upto 0.001mm.

Fig. 3.5 Micro-manometer

3.1.4 Selector switch :- A 24 channel selector switch coupled with a digital display device is used to indicate the temperature at different locations of the test section.

                                      Fig. 3.6 Selector switch

3.1.5 Test section :- The test section consists of copper pipe of length 1000 mm and diameter 25 mm covered with nichrome wire to provide a constant heat flux over the exterier surface of test section. Six thermocouples are tapped over the surface of copper pipe separated by 160 mm distance to measure the surface temperature.The copper tube is insulated by glass wool and covered by black foam sheet.Two thermocouples are placed at inlet and exit of the copper pipe to measure the temperature of water.

Fig.3.7 Test section

3.2 Geometry Of Conical Inserts

During the experimentation, conical inserts of different geometries having same diameter ratio (d/D) of 0.5 and different pitch ratios are inserted into the circular tube to collect the data pertaining to heat transfer and friction. Conical rings are obtained from solid aluminium rod by machining process in lathe machine. The inserts are placed in three different configurations, namely solid conical rings, perforated conical rings and slotted conical rings. The first type of insert as shown in fig. 3.8 is a solid conical ring having length of 23.5 mm, inlet and throat diameter of 23.5 mm and 11.75 mm, respectively. The second type of insert as shown in fig. 3.9 is a slotted conical rings having slot of dimension (15mm '' 1mm) while keeping the other dimensions same as the solid conical rings. The third type of insert as shown in fig. 3.10 is a perforated conical ring having holes of diameter 4.5 mm with the other dimensions same as solid conical rings. All the types of conical inserts are placed at different pitch ratio ( P.R.) of 2, 4, 6 and 8. The thickness of conical rings are kept as 1.5 mm.

Figs 3.10, 3.11 and 3.12  show the different type of conical inserts placed into the test tube.

Table 3.1 Range of conical rings parameters

Conical Rings Solid conical rings Perforated conical    rings Slotted conical rings

Dimension of the rings Length of ring: 23.5 mm Inlet diameter of Conical       ring : 23.5mm

Throat diameter : 11.75 mm          

Thickness of ring : 1.5 mm Same as solid rings Same as solid rings

Pitch length 50 mm, 100 mm, 150 mm, 200 mm Same as solid rings Same as solid rings.

Pitch ratio (P.R.) 2, 4, 6, 8 Same as solid rings Same as solid rings

Material Aluminium Aluminium Aluminium

Number of holes None (n : 2, 4) None

Number of slots None None (s : 2, 4)

Fig. 3.8 Solid Conical Ring

    Fig. 3.9 Perforated Conical Ring

Fig. 3.10 Slotted Conical Ring

Fig. 3.14 Photographic view of solid conical ring insert.

Fig. 3.15 Photographic view of perforated conical ring insert with two holes

           Fig. 3.16 Photographic view of slotted conical ring insert with two slots.

             Fig. 3.17 Photographic view of perforated conical ring insert with four holes.

            Fig. 3.18 Photographic view of slotted conical ring with four slots.'

CHAPTER 4

EXPERIMENTAL DATA COLLECTION

4.1 Experimental Methodology

    The experiments are carried out in a circular tube fitted with conical ring inserts, perforated conical ring inserts and slotted conical ring inserts over a wide range of Reynolds number between 4,000 to 11,000 with water as a working fluid to collect the experimental data related to heat transfer and friction. Experiments are also carried out in a smooth tube to compare the performance of conical inserts. The water from the storage tank is drawn at a  regulated  flow rate by using glove valve and flow rate is measured with the help of rotameter after ensuring the steady state operation. Temperature of water at inlet and outlet as well as test section surface are recorded at different flow rates. Intially the tube surface and the temperature of water at inlet and exit are unstable and approach to higher values with time. After one hour, the temperature readings have negligible variation with respect to time which confirms that the system has reached the steady state condition. After attaining the steady state, the surface temperature at six locations, inlet and exit temperatures of fluid are recorded. The complete process is repeated at different flow rates to collect the data of heat transfer and friction over a range of Reynolds number. The difference in pressure head across the test section is measured with the help of micro- manometer.

4.2 Experimental Data Reduction

 Raw data is used to evaluate the values of dimensionless parameters like Reynolds number , friction factor and Nusselt number. The following method is used to obtain the dimensionless parameters (Reynolds number, friction factor) representing the fluid flow, heat transfer and friction.

The working fluid flows through the heated tube and thus the heat transfer by convection from the heated tube surface is equal to heat received by water, under adiabatic condition.

Qwater  =  Qconvection                                                                                                                                                               (1)

The heat gain by water can be written as

Qwater = m''Cp(Tout - Tin)                                                                                                                   (2)

Where m is mass flow rate of water , Cp is specific heat of water

The convective heat transfer from the heated surface of the tube is equal to

Qconvection = hAs(Ts  - Tb)                                                                                                    (3)

Where As is the surface area of tube, h is the convective heat transfer coefficient, Ts is average temperature of the tube surface, Tb is the average of inlet and outlet temperatures of water

As = ''DL                                                                                                                                           (4)

Where D is inner diameter of the tube and L is length of test section.

Heat transfer coefficient can be obtain by energy balance as:

h = m''Cp(Tout - Tin) / As(Ts - Tb)                                                                                                                                                               (5)

The Reynolds number is calculated as,

Re = U '' D / ''                                                                                                                                   (6)

The Nusselt number is calculated as:

Nu = hD / k                                                                                                                                        (7)

Nu = 5531.68''0.024 / (0.60072)

Nu = 221.002

The friction factor is estimated as:

f = 2D '' 'P / (L''U2)                                                                                                                         (8)

 Where U is average velocity of the fluid and '' is kinematic viscosity of the fluid at temperature Tb.  

For a solid conical ring with pitch ratio of 2, key parameters are found as shown in the following steps:

m'' = 0.222 kg/s

As = 0.07538 m2

From the eqn. (5)

h = 0.222 '' 4187 '' (297.2-296.4) / (0.07538) '' (298.58-296.8)

h = 553.68 W/m2K

Now using the eqn. (6), Reynolds no is determined as:

Re = 0.492 '' 0.024 / (9.8 '' 10-7)

Re = 11,927.51

Where 'P is the pressure drop between the inlet and outlet of the test section and '' is the mass density of fluid at temperature Tb. The mean velocity of fluid (U) is calculated from the mass flow rate of the water as shown below :

U = m'' / ('' '' Ac)                                                                                                                                  (9)

Where Ac  is the cross sectional area of the tube

U =0.263 / (998 '' 0.000452)

U = 0.492 m/s

By using the eqn.(8), the friction factor can be determined as :

'P = 770.68  N/m2

f = 2 '' 0.024 '' 770.68 / ( 998 '' (0.492)2 )

f  = 0.153

The thermo-hydraulic performance factor ('') can be determined as :

'' =( Nu/Nus ) / (f/fs)1/3                                                                                                                                                                          (10)

'' = (221.002/84.254) / (0.153/0.031248)1/3

'' = 1.54

All the thermo-physical properties of the water are calculated at the mean bulk temperature of the fluid (Tb) given by : Tb = (Tin + Tout ) / 2                                                                                 (11)

4.3 VALIDATION OF EXPERIMENTAL SET UP

The experimental data on heat transfer and pressure drop for the smooth tube is collected and verified. The experimental values of the Nusselt number and friction factor of plain tube are compared with the values obtained from standard correlations. The standard data of Nusselt number and friction factor pertaining to flow through smooth tube are obtained from Dittus-Boelter and Blasius correlation [23].

Dittus-Boelter correlation

Nu= 0.023Re0.8Pr0.4                                                                                                                                                            (12)

Blasius correlation

f = 0.316 Re0.25                                                                                                                                                                      (13)

 The experimental values of Nusselt number and friction factor in a smooth tube are compared with the standard correlation data of Nusselt number and friction factor as shown in figs 4.1 and 4.2. The average absolute deviation in Nusselt number and friction factor values are found to be 8.52% and  3.94% which lies within the acceptable limit. .  

Fig. 4.1 Validation of Nusselt number data for smooth tube with the Dittus boelter data.

Fig. 4.2  Validation of friction factor for smooth tube with the Blasius data.

CHAPTER 5

RESULTS AND DISCUSSION

The heat transfer and friction characteristics of a circular tube with conical ring inserts are discussed with the plots of Nusselt number and friction factor as a function of Reynolds number. Effect of inserts on Nusselt number and friction factor are discussed for the entire range of parameters by varying the Reynolds number between 4,000 to 11,000. The heat transfer enhancement at the cost of rising frictional losses is also evaluated by estimating thermo-hydraulic performance for different geometries. The effect of periodically repeated diverging conical ring on the heat transfer and friction factor are discussed with the help of conceptual flow patterns.

5.1 EEFECT OF REYNOLDS NUMBER ON NUSSELT NUMBER

Figures 5.1, 5.2, 5.3 and 5.4 shows the variation of  Nusselt number with Reynolds number for a smooth circular tube with different types of  conical insert geometries in which the diameter ratio is fixed (d/D = 0.5)  and the pitch ratio ( P/D) is varied from 2 - 8. The Nusselt number approached to the maximum value for conical inserts followed by perforated conical inserts and the slotted conical inserts. It can be noticed that Reynolds number has greater effect on Nusselt number as the number of cones are increased in all cases. It may be due to the fact that greater number of cones create more disturbance in the tube and thus extract more thermal energy from the tube surface. The higher values of Nusselt number may be attributed to the disruption in the boundary layer near the wall and thereby proper mixing of the fluid takes place. It can be noticed that in case of perforated conical rings the heat transfer is less as compared to solid conical rings because the holes produce jets impinging on the tube surface which develops parabolic velocity profile and a moderate amount of turbulence upstream. While the heat transfer in slots occur least as compared to perforated conical rings because the geometry of the slots is thin and flat which create an initial flow with a fairly flat velocity profile, less turbulence and a downstream flow  .

Fig 5.1 Variation of Nusselt number with Reynolds number at pitch ratio of 2

Fig 5.2 Variation of Nusselt number with Reynolds number at pitch ratio of 4

Fig 5.3 Variation of Nusselt number with Reynolds number at pitch ratio of 6

Fig 5.4 Variation of Nusselt number with Reynolds number at pitch ratio of 8

5.2 EFFECT OF REYNOLDS NUMBER ON FRICTION FACTOR

 Figures 5.5, 5.6, 5.7, 5.8 show the variation of friction factor with Reynolds number for solid conical rings, perforated conical rings and slotted conical rings at pitch ratio (P.R.) of 2 - 8. It is observed from the results that the friction factor decreases with Reynolds number. At pitch ratio of 2 the friction factor approaches to maximum values for solid conical rings, followed by perforated conical rings and least for slotted conical rings. The heat transfer performance of solid conical rings surpasses the remaining configuration of conical rings irrespective of the Reynolds number. As the pitch ratio is decreasing, value of friction factor is increasing because more number of conical rings are available to create disturbance in the flow. Friction factor is decreasing with Reynolds number due to dissipation of dynamic pressure of the fluid because of high viscous loses near the pipe wall and due to the forces exerted by the ring blockage of the cones. At pitch ratio of 2, friction factor for solid conical rings is 0.27 to 0.15, for perforated conical rings friction factor is 0.21 to 0.13, and for slotted conical rings friction factor is 0.13 to 0.10.

Fig.5.5 Variation of friction factor with Reynolds number at pitch ratio of 2

Fig. 5.6 Variation of friction factor with Reynolds number at pitch ratio of 4

Fig.5.7  Variation of friction factor with Reynolds number at pitch ratio of 6

Fig.5.8 Variation of friction factor with Reynolds number at pitch ratio of 8

5.3 EFFECT OF REYNOLDS NUMBER ON NUSSELT NUMBER ENHANCEMENT

Figures 5.9, 5.10, 5.11 and 5.12 show the variation of Nusselt number enhancement with Reynolds number. The Nusselt number enhancement ratio declines with the increase in Reynolds number in all types of conical inserts. It depicts that the system should be operated in low Reynolds number range to obtain better heat transfer rates.The Nusselt number enhancement ratio is maximum for solid conical ring, followed by perforated conical ring and least for slotted conical ring. The Nusselt  number enhancement ratio for solid conical rings is 4.17, for perforated conical ring is 3.87, for slotted ring is 3.02.

 Fig.5.9 Variation of Nusselt number enhancement ratio with respect to Reynolds number at pitch ratio of 2.

 Fig. 5.10 Variation of Nusselt number enhancement ratio with respect to Reynolds at pitch ratio of 4

Fig. 5.11 Variation of Nusselt number enhancement ratio with respect to Reynolds number at pitch ratio of 6

 Fig. 5.12 Variation of Nusselt number enhancement ratio with respect to Reynolds number at pitch ratio of 8

5.4 EEFECT OF REYNOLDS NUMBER ON FRICTION FACTOR ENHANCEMENT RATIO

 Figures 5.13, 5.14, 5.15 and 5.16 show the variation of friction factor enhancement ratio with respect to Reynolds number. It can be observed from graph that friction factor enhancement ratio decreases as the Reynolds number increases. The friction factor enhancement is maximum for solid cones because more turbulence is created by solid conical rings, then for perforated cones and least for slotted cones. At pitch ratio of 2, the value of friction enhancement ratio for solid cones is 6.95, for perforated cones is 5.48 and for slotted cones is 3.52.

Fig. 5.13 Variation of friction factor enhancement ratio with Reynolds number at pitch ratio of 2.

 Fig. 5.14 Variation of  friction factor enhancement ratio with Respect to Reynolds number at pitch ratio of 4.

 Fig. 5.15 Variation of friction factor enhancement ratio with respect to Reynolds number at pitch ratio of 6.

 

Fig. 5.16 Variation of friction factor enhancement ratio with respect to Reynolds number at pitch ratio of 8.

5.5 EEFECT OF REYNOLDS NUMBER ON THERMO-HYDRAULIC PERFORMANCE FACTOR

Figures 5.17, 5.18, 5.19, 5.20 show the variation of thermo-hydraulic performance factor with regard to Reynolds number for solid conical rings, perforated conical rings and slotted conical rings for different pitch ratio (P.R.) of 2-8. From the graph it is found that the thermo hydraulic performance decreases with increasing Reynolds number and it is maximum at lower Reynolds number. It is observed that the turbulator with smallest pitch ratio provide the highest value of thermo-hydraulic performance. At pitch ratio of 2 the maximum value of thermo-hydraulic performance factor  for solid conical ring is 2.18, for perforated conical ring is 2.11 and for slotted conical rings is 1.98. The thermo-hydraulic performance factor of solid conical rings is higher than perforated and slotted conical rings. The thermo-hydraulic performance factor decreases with Reynolds number because the pressure drop increases due to which more pumping energy is required.

Fig. 5.17 Variation of thermal performance factor with respect to Reynolds number at pitch ratio of 2

Fig. 5.18 Variation of thermal performance factor with respect to Reynolds number at pitch ratio of 4

 Fig. 5.19 Variation of thermal performance factor with respect to Reynolds number at pitch ratio of 6

Fig. 5.20 Variation of  thermal performance factor with Reynolds number at pitch ratio of 8.

It is visible from the discussion in the literature that the conical rings provide better mixing of fluid and greater convective heat flux. Improvement in the heat transfer is caused by the alteration in the fluid flow behaviour near the surface of the tube therefore it is important to reveal the mechanism of heat transfer in cones. The flow pattern in cones presented by Promvonge [12 ] are very helpful to understand the flow mechanism. The flow pattern shown in fig. 5.21 to 5.23 are useful in knowing the flow pattern in cones.

Turbulators or inserts are widely employed in heat transfer engineering applications. The reverse flow is sometimes called 're-circulation flow'. The effect of reverse flow and boundary layer disruption are to enhance the heat transfer coefficient and momentum transfers. The reverse flow with high turbulent flow can improve convection to the tube wall by increasing effective axial Reynolds number.

It is observed that the use of conical ring inserts leads to considerably higher heat transfer rates than plain tube because the centrifugal forces acting on the fluid diverge the flow from central core region towards the wall where high velocity flow impinges over the tube surface, thus extracting greater amount of heat energy. Higher re-circulation and higher contact surface area between the fluid and heating wall extracts more thermal energy.

Perforated conical rings performs better than slotted conical rings because the kinetic energy of fluid which is ejected from the holes is more and the direction of fluid is in upward direction which helps in breaking thermal boundary layer there by causing better mixing between wall and core region of the tube thus enhancing the convective process.

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