INTRODUCTION:

Refrigeration is one of the most important study fields of thermodynamics which is a process to control the heat transportation. Even though we can count on many advantages of refrigeration, this process consequences some environmental problems such as ozone depletion and global warming due to CFC refrigerants. As a result, the researchers tend to concentrate on other non-conventional systems, one of them is a vortex tube which is a mechanical device where anatural substance such as air is used as working medium to achieve refrigeration. It was first discovered in 1933 by the French physicist Georges J. Ranque, and in 1947 the German engineer Rudolf Hilsch improved and modified the design performed by Ranque[1]. After that, the vortex tube was known as the RanqueHilsch vortex tube (RHVT). The RHVT is an effective energy separation device with a simple geometry and has no moving parts. RHVT consists of many parts including one or more inlet nozzles, a vortex chamber, a control valve or a plug that is located at the hot end, a cold end orifice and a working tube. As benefits of using air in RHVT, one can consider low-cost, lightweight, no electricity, reliability, compactness, free maintenance (no moving parts) and without using any refrigerant or electricity having an adjustable temperature range and cools. Other benefits of vortex cooling are fast cooling, short time producing a low-temperature stream, and no moving parts thus thelittle requirement for maintenance.

Using compressed air or any gas as aninput, RHVT tube separates a high-pressure gas into two fluids with distinctly different temperatures, hot and cold fluids. A circular movement inside the vortex tube occurs due to the structure of the tube, depending on its inlet pressure and speed. This movement yields the pressure difference between tube wall and tube center consequences of the friction of the fluid circling at high speeds. It can be observed that the speed of the fluid close to the tube wall is lower than the center one due to the wall friction. Depending on the geometric structure of the vortex tube, energy transfers from fluid in the center region to the tube wall. From the cold region of thevortex tube, the cool fluid leaves and after the stagnation point, the motion is the opposite direction of the main flow. Nevertheless, the heated fluid leaves the tube in the main flow direction from the other end of the tube.

Compressed air enters through the nozzle. At entry, the air expands and acquires high velocity due to theparticular shape of the nozzle. A vortex flow is created in the chamber and air travels in spiral-like motion along the periphery of the hot side at around million rpm. This swirling flow is restricted by the valve. When the pressure of the air near valve is made more than outside by partly closing the valve, a reversed axial flow through the core of the hot side starts from ahigh-pressure region to low-pressure region. During this process, heat transfer takes place between thereversed stream and forward stream. Therefore, air stream through the core gets cooled, while air stream in forwarding direction gets heated up. The cold stream is escaped through the diaphragm hole into the cold side, while hot steam escapes through the valve opening at thehot end. By controlling the opening of the valve, the quantity of the cold air and its temperature can be varied. [2] The cold air thus obtained can be used in applications like air suits for cold mine workers, spot cooling of machining operations, electronic panel and circuit cooling, for setting hot melts, cooling of soldered parts in thecircuit. Thus vortex tube can become agood alternative for commercial vapor compression and absorption refrigeration system that utilizes refrigerants for producing refrigeration effect. Wide-scale commercialization of this concept is still not unveiled because of its low thermal efficiency and inability to produce alarge mass of cold air. Singh et al. [3] varied L/D ratio of the vortex tube and observed that the effect of nozzle design is more important than the cold orifice design for getting higher temperature drops. Cold mass fraction, as well as adiabatic efficiency, is more influenced by the size of the cold orifice rather than the size of the nozzle. GAO [4] with various L/D ratios found that with the length the temperature drop increases. Varying the shape of hot end plug does not affect the performance of the tube. The effect of working tube radius on vortex tube performance and the optimum working tube radius has been determined by Rafiee[5]

Divide and Pise[6] studied the effect of varying conical valve angles on theefficiency of vortex tubes and concluded that for short straight divergent tubes of L/D equal to 8 valve angle of 45º gives maximum cooling effect and valves of 60º angle provide thebetter heating effect. [6]Markal et al. [7] studied the effect of the valve angle of counter-flow Ranque–Hilsch vortex tubes on thermal energy separation, it is observed that the valve angle has a weak influence on the system performance. Seyed and Sadeghiazad[5], carried out an experimental study about thecharacteristic analysis of the performance of a counterflow RHVT with regard to control valve angle. Arzomandi and xue[8] experimented with vortex angle entries the vortex angle plays important role in both the temperature separation and the vortex tube performance. Smaller vortex angle gives larger temperature difference. The above work has identified that vortex tube for low temperature and high cold mass fraction both are different issues and cannot be combined, but combined high CMF and theLow temperature is the need of today, and hence this attempt is made to attain both. Also, the conical valve angle has no effect is stated, the same is proposed to study. K Dincer et al. [9] experimented with conical valves with angles 30°, 60°, 90°, 120°, 150°and 180° and concluded that the biggest temperature difference value of 51°C is observed with the plug which has a tip angle of 30° or 60°.

Vaidyanathan and Prabhakaran [10] concluded from his experiments that when the diameter of the orifice is 6 mm (0.5 D) out of orifice diameter 5 mm, 6 mm and 7 mm, it produces amaximum cold air temperature reduction of 26.5°C.Kırmacı [11], examined the influence of orifice nozzle number and inlet pressure on heating and cooling performance of vortex tube using air and oxygen as test fluids. His results indicated that the temperature difference between the hot and cold fluid decreases with increasing nozzle number.Chang et al. [12] carried out experiments to investigate the influence of divergence angle on the performance of vortex tube and states that the performance of vortex tube can be improved by using a divergent hot tube.Promvonge[13] studied the effect of a number of tangential entries, cold orifice diameter and tube insulation the results show that the insulated vortex tube with 4 inlet nozzles and acold orifice diameter of 0.5D yielded the highest temperature reduction (temperature separation) and isentropic efficiency at about 30ºC and 33% respectively. The increase in the number of inlet nozzles led to higher temperature separation in the vortex tube. Divide and Pise[14] perform anexperiment on inlet pressure (200–600 kPa) and cold mass fraction (0–1) is varied and effect of Mach number at the inlet of the tube is investigated. It is observed that higher COP and low cold end temperature is obtained at subsonic Ma. As CMF increases, COP rises and cold and temperature drops.

Dincer et al. [15] performed experiments with hotplug located at one end, effects of position, diameter (5, 6, 7, 8 mm) and angle (30º–180º) of a mobile plug. The maximum difference in the temperatures of hot and cold streams was obtained for the plug diameter of 5 mm, tip angles of 30º and 60 º, 4 nozzles and by keeping the plug located at the far extreme end. Nimbalkar and Muller [16] investigated for the optimum geometry for cold end orifice of the tube, for different inlet pressures and cold fractions. The experimental results indicate that there is an optimum diameter of cold end orifice for achieving maximum energy separation. Aydin et al. [17] a new geometry is introduced for the cold end side (i.e. where the swirl flow is introduced into the tube), which is called ‘helical swirl flow generator’. S. Eiamsa-and[18] experimented with multiple snail entries to vortex tube. The experimental results reveal that the RHVT with the snail entry provides greater cold air temperature reduction and cooling efficiency. The increase in the nozzle number and the supply pressure leads to the rise of the swirl/vortex intensity and thus the energy separation in the tube.

In this experimentation, the divergent angle 0° and 4° were chosen and L/D is selected in close range 15 to 18 with 2, 4 and 6 inlet nozzle. Orifice diameter ranges from 5 to 7 mm and conical valve angle of 30°, 45°, 60°, 75° and 90°. Pressure range for theexperiment was 2 to 6 bar.

EXPERIMENTAL METHOD:

A. Experimental Setup

The Block diagram of Experimental setup is shown in Fig.1. It comprises of air compressor, pressure regulator, rotameter, air splitter and vortex tube. The pressure regulator was used to regulate the pressure of inlet air and rotameters were used to measure the mass flow rate of theinlet and cold air. RTD thermocouples were used for measurement of thetemperature of cold, hot and inlet air with digital temperature indicator.

Figure 1Block Diagram of Experimental Setup

B. Experimental Procedure

The required amount of compress air was released in the experimental setup via apressure regulating valve and went to rotameter where mass flow rate was measured. The compressed air goes to air splitter where it gets divided into 4 or 6 channel as per requirement. Then it goes to vortex tube either by 2 nozzle inlet, 4 nozzle inlet or 6 nozzle inlet as per experimentation. The conical valve angle at hot end side was adjusted with the help of screw arrangement to get lowest cold air temperature at cold end side. The temperature was measured using RTD sensor. The air temperature was measured at cold end side, hot end side and at ambient. The mass flow rate of cold air was measured with another rotameter attached downstream of the setup.Experiment was conducted on vortex tube having divergent angle of 0 and 4 degree, inlet nozzle of 2, 4 and 6, L/D ratio of 15, 16, 17 and 18, Orifice diameter of 5mm, 6mm and 7mm and conical valve angle (Ɵ) of 30°, 45°, 60°, 75° and 90°.

C. Data Reduction

1. Cold Mass fraction (CMF)

The cold mass fraction is one of the important parameters since it indicates the vortex tube performance and the energy separation inside the vortex tube. Cold mass fraction is defined as the percentage of the pressurized air input to that released through the cold end of the tube and can be calculated by using Eq. 1.The cold mass fraction can be controlled by the conical valve, which is placed at the hot end of the tube. This can be expressed as follows

(1)

2. Specific Heat of Air at Constant Volume

The ideal gas law and first law gives the Eq. 2, which been used to calculate the value of C_V

(2)

3. Adiabatic Gas Constant (γ)

The ratio of the specific heats is a factor in determining the speed of sound in a gas and other adiabatic processes as well as this application to heat engines. This ratio γ is 1.66 for an ideal monoatomic gas and γ is 1.4 for air, which is predominantly a diatomic gas.

(3)

4. Temperature of Compressor (γ)

The temperature at compressor was calculated by Eq. 4

(4)

5. Refrigeration Effect (R_E)

The cooling effect which is produced by a machine is known as refrigeration effect.

(5)

6. Compressor Work (W_comp)

Compressor work was calculated by Eq. 6

(6)

7. Coefficient of Performance (COP)

The ratio of the heat energy extracted by the heat engine at the low temperature to the work supplied to operate the cycle.

(7)

8. Temperature difference (∆T_c)

Temperature difference was calculated by Eq.8

(8)

Design of Simulation Model

The results of CFD analysis and experimental study are summarized and relative comparison has been presented. CFD modelwas developed to calculate the COP and ∆TC at different pressures for adifferent condition. The experimental results presented herein are verified and compared with CFD results. Ansys 15.0 workbench fluent was used to simulate the vortex tube. The geometry, meshing and static temperature distribution were explained below.

A. Geometry

Geometry was made using geometry tools available in theworkbench. Three different geometries of vortex tube were made for 2, 4 and 6 nozzle inlet. The geometry of 2, 4 and 6 nozzle inlet is shown in Fig. 2, 4 and 6. The summary of geometry is given in Table 1.

B. Meshing

As observed from Fig.5the value of cold end temperature remainssignificantly unchanged after grid size is increased beyond582311 cells. And for cells size 371209, the T_C experimental was observed to be near to T_C simulated.

Figure 5. Grid Independence Test based on T_c

The Mesh cell in Fluid Flow analysis systems or the Mesh component system is used to produce a mesh using the Meshing application. It can also be used to import an existing mesh file. When the geometry is imported it must be specified to definite nodes and elements. The meshing of the geometric model of vortex tube contains nodes and elements which specified in Table 2. Below Fig. 6, 7 and 8show the meshing of vortex tube of 2, 4 and 6 nozzle inlet respectively.

Table 2Meshing Summary

Sr. No	Vortex Tube with number of nozzle inlet	Number of Nodes	Number of Elements
1	2	74587	370436
2	4	77693	374248
3	6	74486	371209

Figure 6Vortex Tube Meshing for N = 2 Figure 7Vortex Tube Meshing for N = 4

Figure 8Vortex Tube Meshing for N = 6

RESULT AND DISCUSSION

C. Experimental Result

For obtaining anequation to compare the effect of nozzle number on CMF,COP, and ΔT_C, scatter graph was plotted irrespective of orifice diameter, L/D ratio, valve angle, and pressure.

Figure 9CMF vs. COP for N = 2

Figure 10CMF vs. ΔTc for N = 2

Eq. 9 and 10 was obtain for CMF vs. COP and CMF vs. ΔT_cfor nozzle number 2 from Fig. 9 and 10 respectively.

	(9)
	(10)

Figure 11CMF vs. COP for N = 4

Figure 12CMF vs. ΔTc for N = 4

Similarly, Eq. 11 and 12were obtained for CMF vs. COP and CMF vs. ΔT_cfor nozzle number 3 from Fig. 11 and 12 respectively.

	(11)
	(12)

Figure 13CMF vs. COP for N = 6

Figure 14CMF vs. ΔTc for N = 6

Eq. 13 and 14 was obtain for CMF vs. COP and CMF vs. ΔT_c for nozzle number 6 from Fig. 13 and 14 respectively.

	(13)
	(14)

Fig. 15 plots CMF vs. COP graph for Eq. 9, 11 and 13 with CMF value ranging from 0 to 1 with astep height of 0.05. The COP of nozzle inlet 2 increases with increase in CMF value about 0.1006 to 0.1237. Whereas the COP of 4 and 6 nozzle inlet decreases from 0.2064 to 0.0842 and 0.0971 to 0.071 respectively. This suggests that 2 nozzle inlet perform best and can produce 0.1237 COP for CMF = 1.

Figure 15CMF vs. COP for different N

Fig. 16 plots CMF vs. ΔT_c graph for Eq. 10, 12 and 14 with CMF value ranging from 0 to 1 with astep height of 0.05. The ΔT_cvalue for 2 nozzle inlet increases with increase in thevalue of CMF about 10.1864 to 20.3794. Whereas the value of ΔT_c for 4 nozzle inlet remains nearly same about 16.474 to 17.1614. The ΔT_c for 6 nozzle inlet decreases from 27.47 to 19.842 as CMF rise from 0 to 1. Hence best result was obtained for 2 nozzle inlet which harvests maximum ΔT_c about 20.3794 at CMF = 1.

Figure 16CMF vs. ΔT_C for different N

D. Validation

The result of experimental COP (COP_E) and simulation COP (COP_s) was plotted against Pressure (P_in) in Fig. 17, 18 and 19 for 2, 4 and 6 nozzle number respectively. From Fig. 17 which was plotted for nozzle number 2, it seems that the value of COP_Eincreases from 0.055083794to 0.136668811 with anincrease in P_in from 2 to 6. The value of COP_Sincreases from 0.046986215to 0.062378066with anincrease in P_in from 2 to 6. The average variation between the COP_EandCOP_Swas0.044821248.

Figure 17Pressure vs. COP Comparison for Simulated and Experimental Value for N = 2

Figure 18 was plotted for nozzle number 4. The value of COP_Eincreases from 0.021 to 0.041 with anincrease in P_in from 2 to 6. The value of COP_Sincreases from 0.018456415 to 0.023497025 with anincrease in P_in from 2 to 6. The average variation between the COP_EandCOP_Swas0.009804236.

Figure 18Pressure vs. COP Comparison for Simulated and Experimental Value for N = 4

Figure 19 was plotted for nozzle number 6. The value of COP_Eincreases from 0.001112469 to 0.025761167 with anincrease in P_in from 2 to 6. The value of COP_Sincreases from 0.00069969 to 0.008031716 with anincrease in P_in from 2 to 6. The average variation between the COP_EandCOP_Swas0.007692762.

Figure 19Pressure vs. COP Comparison for Simulated and Experimental Value for N = 6

The result of experimental ∆T_C (∆T_CE) and simulation ∆T_C (∆T_CS) was plotted against Pressure (P_in) in Fig. 20, 21 and 22 for 2, 4 and 6 nozzle number respectively. From Fig. 20 which was plotted for nozzle number 2, it seems that the value of ∆T_CEincreases from 2.5 to 7.2 with anincrease in P_in from 2 to 6. The value of ∆T_CSincreases from 3.60001 to 11.80075 with anincrease in P_in from 2 to 6. The average variation between the ∆T_CE and∆T_CSwas3.08979.

Figure 20Pressure vs. ΔT_C Comparison for Simulated and Experimental Value for N = 4

Figure 21 was plotted for nozzle number 4. The value of ∆T_CEincreases from 11.7to 17.92 with anincrease in P_in from 2 to 6. The value of ∆T_CSincreases from 19.2 to 29.4 with anincrease in P_in from 2 to 6. The average variation between the ∆T_CEand∆T_CSwas 6.56.

Figure 21Pressure vs. ΔT_C Comparison for Simulated and Experimental Value for N = 4

Figure 22 was plotted for nozzle number 4. The value of ∆T_CEincreases from 3.5to 16.7 with anincrease in P_in from 2 to 6. The value of ∆T_CSincreases from 2.79404 to 13.51755 with anincrease in P_in from 2 to 6. The average variation between the ∆T_CEand∆T_CSwas1.911164.

Figure 22Pressure vs. ΔT_C Comparison for Simulated and Experimental Value for N = 6

Static temperature is atemperature with no velocity effect. Static temperature distribution was studied in vortex tube to determine the temperature separation phenomena occurring in a different number of inlet nozzle. Fig. 23, 24 and 25show static temperature distribution of vortex tube at apressure of 6 bar for different inlet nozzle number.

Figure 23Static Temperature Distribution of Vortex Tube for N = 2

From Fig. 23 which is vortex tube with nozzle number 2, it could be seen that the temperature variation profile is developing along the length of vortex tube on both sides. The cold end side temperature which is lower than the atmospheric temperature is clearly visible and also the hot end side temperature is higher than the atmospheric temperature.

Figure 24Static Temperature Distribution of Vortex Tube for N = 4

From Fig. 24 which is for nozzle number 4, the temperature variation profile is also developing on both way along the vortex tube length but the region of the temperature variation profile generated is lesser than the region of the temperature variation profile developed in nozzle number 2. The cold end temperature is lesser than atmospheric temperature but greater than the cold end temperature achieved for nozzle number 2.

Figure 25Static Temperature Distribution of Vortex Tube for N = 6

Fig. 25 is for nozzle number6. The temperature variation profile developed only enhance in thecold end side of vortex tube along its length. This suggests that maximum compress air is escaping from cold end side without generating any significant temperature variation profile. The cold end side temperature achieved is also more as compared to nozzle number 2 and 4. And hot end side temperature is less as compared to the result obtained for nozzle number 2 and 4.It could be suggested that the greater the temperature variation profile developed inside the vortex tube, the better will be temperature separation and thebetter cooling effect will be achieved.

CONCLUSION

The following are the conclusion of experimentation and simulation:

1. The COP of nozzle inlet 2 increases with increase in CMF value about 0.1006 to 0.1237. Whereas the COP of 4 and 6 nozzle inlet decreases from 0.2064 to 0.0842 and 0.0971 to 0.071 respectively. This suggests that 2 nozzle inlet perform best and can produce 0.1237 COP for CMF = 1.

2. The ΔT_C value for 2 nozzle inlet increases with increase in thevalue of CMF about 10.1864 to 20.3794. Whereas the value of ΔT_C for 4 nozzle inlet remains nearly same about 16.474 to 17.1614. The ΔT_C for 6 nozzle inlet decreases from 27.47 to 19.842 as CMF rise from 0 to 1. Hence best result was obtained for 2 nozzle inlet which harvests maximum ΔT_C about 20.3794 at CMF = 1.

3. The maximum COP_EandCOP_S value 0.136668811 and 0.0623780366 respectively was achieved at 6 bar pressure for 2 nozzle inlet.

4. The maximum ∆T_CEand ∆T_CS value 7.2 and11.80075 respectively were achieved at 6 bar pressure for 2 nozzle inlet.

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Received on 18.09.2017 Accepted on 15.01.2018

Int. J. Tech. 2018; 8(2): 65-78.

DOI:10.5958/2231-3915.2018.00010.X

Measurement	Present Vortex Tube
Length	187.5 mm
Diameter	12.5 mm
Orifice diameter	6 mm
Valve Angle	45º
Nozzle Diameter	4 mm
Number of Nozzles	2, 4 and 6
Diverging Angle	0º