SECTION 8 HEAT TRANSFER - UAB · SECTION 8 . HEAT TRANSFER . ... coefficient, whereas as the ... spent fluid flow and the impact of cross flow between the jet

SECTION 8

HEAT TRANSFER

ASME District F - ECTC 2013 Proceedings - Vol. 12 257


ASME District F - Early Career Technical Conference Proceedings ASME District F - Early Career Technical Conference, ASME District F – ECTC 2013

November 2 – 3, 2013 - Birmingham, Alabama USA

INVESTIGATION INTO THE THERMAL HOMOGENEITY OF JET IMPINGEMENT HEAT TRANSFER IN CONCENTRIC TUBE ANNULUS

Kennedy Osaighe Amedu

Mechanical Engineering Department, Faculty of Engineering and Environment, Northumbria University Newcastle upon Tyne, Tyne and Wear, United Kingdom.

ABSTRACT In order to evaluate the problem of non-thermal

homogeneity during the cooling of a plunger in the narrow neck press and blow process in the glass industry, investigation has been carried out to evaluate the heat transfer performance of the cooling tube of a concentric tube heat exchanger. Two experiments were carried out after a theoretical design of an inline jet array with jet height to diameter (𝐻/𝐷) of 4.5, having central jet 𝐻𝑐/𝐷𝑐 of 5 and jet to jet spacing s of 4.9𝑚𝑚 (𝑆/𝐻 of 1.4), forming a 6x4 configuration and an additional 5x4 configuration designed on the 6x4 configuration were carried out.

Six Reynolds numbers, Re = 2845, 8032, 9539, 11881, 13722 and 15898 of the working fluid were also investigated. Detailed local temperature and Nusselt number distribution were analysed for both cases. The average Nusselt number was also derived and compared. Data analysis shows that the inline jet array with equal pitch of 4.9𝑚𝑚 (6x4 configuration) produced good thermal homogeneity and higher Nusselt number compared to the 5x4 extra configuration designed on the 6x4. Also experimental results were compared with available correlation equations for an array of jets, and the prediction was observed to be good. A methodology for designing the cooling tube in order to achieve thermal homogeneity was proposed.

INTRODUCTION Impinging jets have been given much attention in industry

and academia because of their impressive high rate of heat and mass transfer either to cause cooling, heating or drying [15]. Apart from areas such as cooling of turbine blades, drying of paper, cooling of electronic devices and the de-icing of aircraft systems [9, 24], there is a critical application in the glass industry during the molten glass gob pressing process in which a plunger is cooled internally with impinging air jets via a concentrically located perforated conical tube as it presses the molten glass to form a hollow glass article in the narrow neck press and blow process. As the plunger travels down the high temperature molten glass, which is at about 1150oC entrapped in the mould, its temperature increases due to conduction of heat from the molten glass leading to variable temperature distribution across the plunger wall. Observations show that the

tip of the plunger experiences the highest surface temperature, since it has more contact with the hot glass. Thus the essence of the cooling medium which is employed in the internal walls of the plunger to moderate the temperature of the molten glass which should be kept just around the adhesion temperature in order to avoid sticking to the plunger surface. Because of the uneven temperature distribution on the plunger wall and the limited understanding of the processes occurring during the cooling of the plunger, it is difficult to achieve uniform cooling in the internal walls of the plunger, which then transcends to the hot glass, thereby affecting the quality of the final glass product. Thus it is of great importance to properly design the cooling tube for effective cooling so as to achieve near thermal homogeneity across the plunger walls.

An impinging jet is a high velocity, usually turbulent jet of fluid mass, which impinges normally on a surface for enhanced coefficients of convective cooling, heating or drying [15]. There are three major characteristic zones (fig. 1) of the flow field, which are: the mixing zone, the impingement zone and the wall jet zone [16].

In fig. 1 below, 𝐿 represents the jet nozzle-to-impingement surface height 𝐻, 𝐷 is the nozzle or jet diameter, 𝑧 is the distance upward from the impingement surface and 𝑟 is the radial distance from the point of stagnation of the impingement surface.

Figure 1 – Flow field of an Impinging Jet: (1) Mixing zone, (2) Established zone, (3) Impingement zone, (4) Wall jet

zone (source: [26]).

Comprehensive reviews have been carried out on confined jet heat transfer [16, 19, 31]. In [1, 12], the heat transfer characteristics of an asymmetric round jet impingement were experimentally studied. The jet Reynolds number, jet to impingement surface distance 𝐻 and the nozzle diameter 𝐷


where given as the basic factors that influence the single confined jet impingement on a flat plate.

In a confined multiple jet array for impingement heat transfer, particularly for asymmetric geometry, an important consideration is the geometrical arrangement of the jets. For design considerations, the impingement heat transfer of a confined asymmetric jet array is a function of the jet to jet spacing (pitch), the jet to impingement surface height (𝐻), the exhaust location and geometry and the Reynolds number of the flow [29]. An important review of the heat transfer in impinging jets was carried out by [26]. Turbulent mixing in the free jet zone, formation of boundary layer, recirculation, physical behaviour of jet impingement and the experimental methods as well as the techniques for measuring heat transfer coefficients for both single and multiple circular and slot jets were discussed. Factors such as turbulence in the free jet, formation of boundary layer, stagnation and recirculation affect the heat transfer coefficients of single impinging jet with multiple jet impingement also exhibiting similar characteristics.

But for real industrial applications, three possible differences, peculiar to multiple jet impingement, occur based on the geometric conditions which are: jet interaction in the mixing zone before impingement, which is dependent on jet to jet spacing; lateral interaction of adjacent jets after impingement, which can possibly lead to significant jet fountains; jet to jet interaction when the geometry is asymmetric and there is no exhaust ports between jets, in which case the spent air interacts with the impinging jets as it flows towards the exhaust location leading to what is referred to as a cross flow [26, 32]. It was also shown that jet interaction can lead to a reduction in the energy of the jet. It stated that the effect of the interaction of the jets when there is no cross flow is dependent on the non-dimensional quantities: jet spacing 𝑆/𝐻 and 𝐻/𝐷 ratios, where 𝑆 is the pitch or distance between two jets. It further elaborated that a large 𝑆/𝐻 ratio reduces fountain interaction and the strength of impingement.

An experimental investigation on confined circular air jets in a staggered format impinging vertically on a flat plate was done by [24], varying the Reynolds number, jet spacing and height and derived an optimum correlation for the jet spacing to the diameter ratio to achieve the maximum heat transfer coefficient in the stagnation zone. Observations showed that jet interference leads to the reduction of the strength of the jets which eventually results in lowering the total heat transfer whereas jet fountain produced a flow recirculation between the fountain and a centre jet considered in the investigation. The investigation also showed that the effect of entrainment leads to a reduction in the heat transfer of the jet array. An extensive investigation into jet array was also done by [19]. It was observed that the results for the local heat transfer coefficients were in close agreement with that of a single jet. The difference observed was the presence of secondary stagnation zones as a result of adjacent wall jets impinging on each other. This gave rise to some peaks in the lateral variation of the heat transfer coefficients.

The outlet flow configuration also determines the extent of uniformity of the heat transfer coefficients over the cooled area in a confined system [10, 19]. The heat transfer coefficient will be more uniform if the ratio of outlet flow area to the jet exit area is greater than unity. Continuous movement of the entrained fluid has been observed to yield uniform distribution of the heat transfer coefficients in the 𝑋 direction. This will not be the case for flow in the 𝑌 direction which can lead to unequal distribution of the temperature leading to poor product quality [19]. Also for lateral flow of fluid through a flow tube for impingement heat transfer, the outlet stream may greatly influence the whole flow field which could simultaneously affect the temperature and concentration fields [10, 19]. The effect of the jet array pattern on the turbulent flow characteristics of the outlet hole in short confined channels was also determined. Of great interest in their research was the flow field in the exhaust hole. It was shown that when the aspect ratio (ratio of the height of the flow passage to the exhaust hole diameter) is unity, the jet array has no significant effect on the flow characteristics but played a crucial role on the discharge coefficient, whereas as the aspect ratio increased to 3 and 5, the flow distribution became more uniform which consequently improved the entry condition and the aerodynamic loss in the exhaust hole. An increase in the discharge coefficient 𝐶𝑑 was also found for the high aspect ratio.

Cross flow due to spent fluid has also been shown to affect both the uniformity and rate of heat transfer of jet array impingement in confined systems [19, 30]. Cross flow occurs as a result of spent fluid flowing normal to the impingement flow as it locates the exhaust [30]. In [30], a 3-D transient liquid crystal scheme and hot film measurement of the flow characteristics was used to study the effect of cross flow in confined cavity and observed that though the heat transfer rate is reduced as a result of cross flow, the uniformity is improved. Also, a maximum heat transfer rate could be attained if the spent fluid flow and the impact of cross flow between the jet exit and impingement surface is properly designed [17, 19, 29]. Though cross flow can lead to the reduction of the heat transfer rate, when the jet to jet spacing is very small, the heat transfer performance can be improved [6, 7, 18, 29]. The effect of cross flow which is dependent on the direction of exit of the spent fluid [12, 19, 29], can further be minimised when the exhaust for the spent fluid flow is located on the impingement surface [11, 14]. For both inline and staggered array of jets for impingement heat transfer, the interaction of adjacent jets leads to a radial network on the impingement surface which significantly influences the distributions of the spatial heat transfer coefficient. This radial network is considered as the hexagonal and square web [20, 29]. A secondary heat transfer zone is developed which surrounds the central jet in asymmetric confined array of jets when the jet Reynolds number is large and the jet to jet spacing is small [20, 26, 29]. The secondary heat transfer rate will increase with an increase in Reynolds number and a decrease in the jet to jet spacing [13, 19, 26] and also a decrease in the asymmetric jets to target surface spacing [13, 26].


There are very few studies on jet array impingement heat transfer in a confined concentric tube annulus. This form of geometry is a case of heat transfer on a curved surface, which could either be concave or convex. Some researchers have been able to make comparisons between jet impingement on flat surface and that on a curved surface and stated that the same approach is followed in designing jet array impingement [3, 29]. Two approaches are usually employed in spacing round jets; inline array of jets, which form a square web and the staggered array of jets, which forms a hexagonal web [15, 19]. As stated by [3, 29], in spacing the jets on a curved surface, the square or hexagonal web forms the radial network of jets. The radial network of jets is uniformly spaced with equal pitch. For the square web of jets giving the inline array, the radial network will therefore form an angle of 90 degree in the loop to make it equal spacing, while for the hexagonal web (staggered array), three adjacent jets form an equilateral triangle with each other for the spacing to be equal.

Another factor that comes into play is the angle of impingement of the jet. Basically, jets perform most efficiently in cooling over a stagnation region with a normal impingement [19]. In [4], an investigation of jet array impingement heat transfer in a concentric annular channel with and without rotating inner cylinder was done. They discovered that for the static cylinder case, the heat transfer is affected by the jet to jet interaction both axially and angularly, entrainment of spent fluid flow creating a cross flow and the Reynolds number. It was observed that increases in Reynolds number in the concentric annulus do not considerably increase the heat transfer but rather aids more mixing to inform a uniform distribution of local temperature and Nusselt number respectively. In [3], a detailed heat transfer measurement was carried out over a convex-dimpled and smooth surface of impinging jet array with three eccentricities including an eccentricity of zero, varying the Reynolds number between 5000 ≤ 𝑅𝑒 ≤ 15000 and jet pitch to diameter ratio (𝑆/𝐷) between 0.5 ≤ 𝑆/𝐷 ≤ 11. The findings by [4] was corroborated, stating that heat transfer in the concentric tube annulus is strongly affected by the jet to impingement surface spacing, jet pitch, Reynolds number and the manner with which the spent flow is treated. The way the spent flow is treated was shown to be important in achieving uniformity of heat transfer over the impinging surface. It was observed that when the array of jets impinging on a surface is confined, the spent fluid could generate a cross flow that diffuses the jet momentum.

In this investigation, a confined concentric inline array of impinging circular air jets drilled on a single nozzle with a central (exit) jet in the form of a concentric tube for forced convection purposes is considered. The spent fluid is made to flow laterally concentrically over the impingement surface (i.e. on both sides of the impingement surface in asymmetrical form. Thus the exhaust is located on the impingement surface housing the concentric cooling tube. Of specific interest is the determination of the jet to target spacing, jet to jet spacing and Reynolds number that would enable the achievement of uniform coefficient of heat transfer distribution on the wall of

the impingement surface. The flow is turbulent and a range of turbulent Reynolds numbers that are applicable in an industrial setting is considered.

EXPERIMENTAL INVESTIGATION For experimental purposes, a prototype concentric tube

heat exchanger was designed. As shown in figure 2, it consists of an inner and outer tube of 6.3mm and 13.5mm diameter respectively. The 6.3mm inner tube was concentrically located vertically inside the outer tube for impingement heat transfer, making the horizontal distance of the lateral side of the inner tube from the inner wall of the outer tube to be 3.6mm. Jet holes of 0.8mm diameter (𝐷) were drilled laterally on the inner tube in an inline array pattern. Thus the horizontal distance forms the jet height (𝐻) for the 0.8 mm holes drilled. This gives an 𝐻 𝐷⁄ ratio of approximately 4.5, making sure the horizontal height is within the potential core of the turbulent jet within which the velocity profile is uniform [8]. Based on the horizontal jet height, the 0.8 mm jet holes were spaced to form the pitch (𝑆) of the array of the jet holes. As has been recommended for effective heat transfer rate [19], the pitch 𝑆 of the jets considered was 1.4 times the jet height 𝐻. In [19] it was suggested that a jet to jet spacing of 1.0 to 1.5 of the jet height from the impingement surface was necessary for effective rate of heat transfer with a recommended optimum of 1.4𝐻. Due to experimental constraints, the 1.4𝐻 was adopted. A central (main) jet of the working fluid (air) of 1.5mm diameter was designed. This central jet has a vertical height (𝐻𝑐) of 7.5mm from the end inner wall of the outer tube. The exhaust hole is designed on the impingement surface based on recommendations to minimise cross flow effect and enhance the possibility of obtaining uniform temperature distribution. The length of the inner and outer tubes was 74.8 mm. The outer tube of thin wall and high conductivity will be considered to be at constant temperature.

Figure 2 – Sectioned view of the concentric tube heat exchanger: (1) Inner cooling tube, (2) Outer tube, (3)

0.8mm Lateral Jet, (4) 1.5mm Central Jet, (5) Outer tube wall.

Figure 3 below shows the experimental rig which consist

of the concentric tube heat exchanger test module, a rotameter which measures the flow rate, water bath, which serves as the heating medium and compressed air supply as the cooling fluid.


Figure 3 – Experimental rig.

The diameter ratio of the concentric tube heat exchanger is

approximately 0.5 and an area ratio of 0.2. The heat exchanger material is copper because of its high thermal conductivity. Four rows of circular jet holes of 0.8 mm diameter were drilled on the inner tube with equal angular interval of 90 degree. Each row consists of six jet holes with equal pitch of 4.9mm, in addition to the 1.5mm central jet, making it a total of 25 jets. The exhaust of the spent air is designed on the impingement surface geometry (outer tube) such that cross flow effects are minimised.

Eight K-type thermocouples were available for the experiment. Six of the K-type thermocouples equally spaced was inserted through grooves drilled in the outer tube in such a way that the local jet temperatures 𝑇𝑗 around the passage area of the annular space can be determined while the other two thermocouples measured the temperatures of the inlet and exhaust air respectively. The annular space for the entrained air has an area of approximately 112mm2. The thermocouples are held in place and sealed to avoid leakage of the impinging air jets on the outer tube using high temperature glass filled epoxy. After this was done, more holes were drilled on the cooling tube to increase the flow area. An additional 5x4 matrix of jets were drilled in between two rows of four jets exactly at the centre point of the four jets at an angle of approximately 45 degrees, on the cooling tube in addition to the 6x4 inline array initially drilled. This gave a uniform spacing laterally of the 5x4 matrix of jets but a non-uniform spacing with the 6x4 inline array of jets. This has been done to ascertain the effect of increasing the flow area with a non-uniform spacing.

EVALUATION OF THE HEAT TRANSFER For determining the heat transfer, experimental data is

recorded at steady state. The steady state condition is determined when the differences in temperatures of the wall of the outer tube after successive measurements is within 0.3℃. The convection heat transfer mode is characterised by energy transfer by a combination of conduction and bulk fluid motion. In the present circumstance, cold air is used as the working fluid flowing through the inner tube to cause cooling on the inner walls of the outer tube. The assumptions made in determining the heat transfer coefficient are steady state conditions, uniform surface temperature of the outer tube, isothermal boundary conditions in the outer wall of the outer tube and incompressible fluid flow.

The rate of heat transfer at steady state is assumed to be equal to the rate of heat loss in the experiment. Therefore we can write that:

𝑄𝑎𝑖𝑟 = 𝑄𝑐𝑜𝑛𝑣 (1) Where: 𝑄𝑎𝑖𝑟 = 𝑚𝐶𝑝𝑎(𝑇𝑜 − 𝑇𝑖) (2) 𝑄𝑐𝑜𝑛𝑣 = ℎ𝑎𝐴𝑠(𝑇𝑠 − 𝑇𝑚) (3) 𝑇𝑚 = 𝑇𝑜+𝑇𝑖

2 (4)

Where 𝑇𝑜 and 𝑇𝑖 are the outlet and inlet air temperature respectively, 𝑇𝑠 is the surface temperature of the inner wall of the outer tube, 𝑇𝑚 is the mean temperature of air, 𝐶𝑝𝑎 is the specific heat capacity of air and 𝑚 is the mass flow rate of air at constant pressure which is calculated from the volumetric flow rate of air derived from the rotameter readings. This is determined from the expression given as:

𝑚 = 𝑄𝑣 × 𝜌 (5) where 𝑄𝑣 and 𝜌 are the volumetric flow rate and density of air at the inlet into the test section. The Reynolds number of the air flow in the annulus is calculated from the equation below:

𝑅𝑒 = 𝜌𝑉𝑚𝐷ℎ𝜇

(6) where 𝜇 is the viscosity, 𝑉𝑚 and 𝐷ℎ are the average velocity and hydraulic diameter respectively of air in the annulus and are obtained from the expression:

𝑉𝑚 = 𝑄𝑣𝜋(𝐷𝑜2−𝐷𝑖

2)4

(7)

𝐷ℎ = 4𝐴𝑐𝑃

=4(𝜋4)(𝐷𝑜2−𝐷𝑖

2)

𝜋𝐷𝑜+𝜋𝐷𝑖= 𝐷𝑜 − 𝐷𝑖 (8)

Also ℎ𝑎 gives the average heat transfer coefficient of the configuration. The average Nusselt number 𝑁𝑢𝑎 is then given as:

𝑁𝑢𝑎 = ℎ𝑎𝐷ℎ𝐾

(9) The local Nusselt number 𝑁𝑢𝑜 of the jets in the annular

region is determined from the local heat transfer coefficient ℎ𝑜 which is obtained by determining the average jet temperatures 𝑇𝑗 measured by the thermocouples.

𝑄𝑐𝑜𝑛𝑣 = ℎ𝑜𝐴𝑠�𝑇𝑠 − 𝑇𝑗� (10) 𝐴𝑠 = 𝜋𝐷𝑜𝑙 (11)

where 𝐴𝑠 and 𝑙 are the inner surface area and length of the outer tube respectively. The local Nusselt number can then be obtained from the equation:

𝑁𝑢𝑜 = ℎ𝑜𝐷ℎ𝑘

(12) From the equations above, the heat transfer coefficients and Nusselt number can be determined for individual jet array pitch arrangement and a comparison can be made to determine their performance.

UNCERTAINTY ANALYSIS The quality of experimental results obtained is basically

affected by the measurement equipment such as the rotameter, thermocouples and the data logger. Another aspect of error is in losses due to natural convection and radiation as the heat exchanger outer surface was not insulated since it was immersed in a large mass of water with the assumption that the


temperature is constant. Also variations that could occur in the fluid properties are taken care of by the mean temperature 𝑇𝑚. The losses in the rotameter calibration are likely to be in the range of ±5%. The errors in the thermocouple readings at steady state temperatures are between 0.3− 0.5℃. Colossal losses are expected to be derived from natural convection to the surroundings which could subsequently affect the calculated heat flux.

RESULTS AND DISCUSSION The results obtained from the experiments carried out

based on the design of the concentric tube inline jets with 𝑆/𝐷 of 6 (4.9mm pitch) and angular distance of 90 degree forming a 6x4 matrix of jets with a central main jet of 𝐻𝑐/𝐷𝑐 of 5 (total of 25 jets) and that obtained by drilling a 5x4 matrix of jets at 45 degree in between a square web formed by the radial network of jets in the first experiment (total of 45 jets) are discussed in this section. The effect of the jet to jet interaction, confinement and Reynolds number on the local temperature distribution, coefficient of heat transfer and the local and average Nusselt numbers are analysed.

From past research studies on the flow visualization of impinging jet array, the heat transfer is considered to be influenced by jet interference before impingement and/or jet fountain [25]. When the jet spacing is very small, there will be interference before impingement between adjacent jets as a result of shear layer expansion. It is believed that this interference brings about a weakening of the strength of the jets resulting in a low heat transfer performance. But when the jet pitch is reasonably large, a fountain occurs between adjacent jets. The occurrence of a fountain effect considerably strengthens the jets.

In a confined annular flow situation as pictured in this investigation, the entrainment of the flow fountain from adjacent jets formed after impingement leads to a recirculation back into the potential core making the heated air to mix with the cold air within the potential core of the jets [4, 24]. The mixing of the jets occurs both in the axial and angular directions respectively. There are four jet rows of high momentum each 90 degree apart impinging on the outer tube. As a result of entrainment, the angular and axial wall jet spent fluid interact with the impinging jets leading to an axial increase of the outer tube wall temperatures from the stagnation zone as it flows annularly towards the exhaust [4]. This subsequently reduces the local Nusselt number.

Also, due to collisions of the angular spent fluid flow between adjacent rows of jets in conjunction with jet to jet interaction before and after impingement, there is a further weakening of the momentum fluxes from the angular wall jet flows, which eventually influences a slight increase in the local Nusselt number in between the axial distribution. Thus due to the weakening of the jet momentum, the increase of the temperature of the fluid and the boundary layer thickening in the annulus, the heat transfer performance is expected to be less in the axial locations between adjacent jet rows.

Changes in the flow characteristics and the distributions of the annular wall jet flows due to changes in the jet to jet interferences and the spent fluid can considerably lead to changes in the heat transfer performance in the concentric narrow annulus due to entrainment [4]. This is evident in the average Nusselt number distributions as shown in figure 9. As the Reynolds number increases, more mixing will occur due to the factors described above and the local Nusselt number is seen to increase. The influence of the combination of jet to impingement surface height, jet to jet interaction, cross flow and Reynolds number determine the uniformity of the local temperatures and Nusselt numbers respectively. The results for both experimental designs for six different Reynolds number are presented below.

Table 1 – Annular Reynolds Number 𝑹𝒆 = 𝑹𝒆𝒂

𝑅𝑒1 𝑅𝑒2 𝑅𝑒3 𝑅𝑒4 𝑅𝑒5 𝑅𝑒6 2845 8032 9529 11881 13722 15897

LOCAL TEMPERATURE DISTRIBUTIONS Figure 4 shows the local temperature distributions of the

6x4 matrix of jets for each Reynolds number considered. It can be deduced that the local temperature begin to drop as the Reynolds number increases as shown from 𝑅𝑒1-𝑅𝑒6. Due to the effects described above, it is observed that the mixing of the fluid tend to bring about the uniformity of the local temperatures. This uniformity is seen to be more pronounce for Reynolds number of 11881. As the Reynolds number was increased from 𝑅𝑒4 to 𝑅𝑒5 and 𝑅𝑒6, the local temperatures subsequently drop, but the temperature difference is seen to rise a bit. For each of the results in fig. 4, it is observed that the highest annulus jet temperature occur close to the tip of the outer tube. Depending on the jet pitch considered and the height of the central jet from the tip of the cylinder, this variation could possibly be reduced.

Figure 4 – Axial Local Temperature distribution for inline

array (6x4 matrix jets). Critically examining the local temperatures for each jet as

recorded by the thermocouples, individual local temperature drops as the Reynolds number increases, indicating gradual cooling locally. This ideally should be the expected trend since there is an increase in the velocity of the working fluid. But depending on the level of mixing due to recirculation of the fountain, cross flow as a result of entrainment or confinement


of the spent fluid in the axial and angular directions, the local temperatures will either increase or decrease which gives a quite complicated flow field. The increase or decrease of the local temperatures tends to reduce the level of variation of the local temperatures. This effect is quite strong for Reynolds number 𝑅𝑒4 of 11881 shown in figure 5.

Figure 5 – Local temperature distribution of inline array at

𝑅𝑒 = 11881 Figure 6 below shows the local temperature distribution for

the 45 jets (6x4 plus 5x4 matrix). There is a comparable similarity in the flow structures as observed in fig. 4 looking at the trend. This could likely be because of the flow pattern in a concentric tube annulus with a central jet. It is also interesting to observe that the 6x4 square webs of radial network of jets gave better performance in terms of rate of heat transfer and uniformity of local temperatures and consequently, the local Nusselt numbers than the 6x4 plus 5x4 matrix of jets which has more flow area. This phenomenon could possibly be due to increased effect of cross flow and jet to jet interference before impingement which eventually degenerated the strength of the jet.

Figure 6 – Local temperature distribution (6x4 by 5x4 jets)

Figures 7 and 8 show the local Nusselt number for both

experimental designs described above. Observations show that the local Nusselt numbers do not considerably change as the Reynolds number is increased rather the increment in Reynolds number aid more mixing to influence uniformity due to the nature of the flow structures developed in the concentric annulus (4, 19, 24). The local Nusselt number is axially symmetric about the mid span axis of 𝑋 𝐿⁄ = 0.5 (4).

Figure 7 – Axial local Nusselt number for Inline array (25

jets).

Figure 8 – Axial local Nusselt number (45 jets).

AVERAGE NUSSELT NUMBER Figures 9 (a) and (b) show the average Nusselt number for

both concentric heat exchanger designs in the experiment for the range of Reynolds number 𝑅𝑒1 to 𝑅𝑒6. For the average Nusselt number for the 6x4 matrix of square web of jets in fig. 9a, there was an increase in the Nusselt number from 68 to 171 as the Reynolds number increased from 𝑅𝑒1 to 𝑅𝑒2, giving the first maximum. It then dropped slightly to 168 for 𝑅𝑒3. This drop could be attributed to the increasing cross flow effect experienced in this circumstance. Thus it is reasonable to say that in the absence of cross flow, the average Nusselt number is expected to increase consistently as the Reynolds number increases. The same impact of cross flow of spent fluid is observed for the Reynolds number 𝑅𝑒4, 𝑅𝑒5 and 𝑅𝑒6 giving Nusselt numbers of 236, 248 and 247 respectively.

In figure 9(b) for the 6x4 plus 5x4 matrix of jets, the average Nusselt number is also seen to increase steadily as the Reynolds number increased. But making a comparison with fig 9a, it can be seen that the average Nusselt number for individual Reynolds number is greater for the inline array of 25 jets than for the 45 jets. This poorer performance of the 45 jets with more flow area could also be attributed to the fact that there was higher jet interference before impingement and cross flow effect in the confined concentric annulus which eventually weakens the strength and heat transfer rate of the jets.


9(a)

9(b)

Figure 9 – Effect of Reynolds number on average Nusselt number: (a) Inline array of jets, (b) 45 jets.

COMPARISON OF EXPERIMENTAL RESULTS WITH LITERATURE

Comparison of the spatially averaged Nusselt number in the two experiments carried out is made with the standard correlation given by [19] for a regular (square or hexagonal) array of round nozzles or jets on a surface given as: 𝑁𝑢𝑎 = 𝑃𝑟0.42(0.5𝑅𝑒0.667), valid for 2000 < 𝑅𝑒 < 100,000 where 𝑁𝑢𝑎 is the average Nusselt number, 𝑃𝑟 is the Prandtl number which is 0.72 for air and 𝑅𝑒 is the Reynolds number. The results of the comparison are tabulated in table 2 below.

Table 2 – Comparison of Average Nusselt number for

Correlation with Experimental Results 𝑹𝒆 𝑵𝒖𝒂(Correlation) 𝑵𝒖𝒂(Inline

array) 𝑵𝒖𝒂(45 jets)

2845 86.57 68.02 54.99 8032 172.95 171.40 144.00 9539 194.00 168.96 158.80 11881 224.69 236.07 187.69 13722 247.24 248.40 198.93 15898 272.75 246.78 224.57 In figure 10 below, it can be observed that a good

comparison and similarity between correlation results and experimental results could be made. The inline array is seen to predict the correlation results more than the matrixes of 45 jets.

Figure 10 – Comparison of Experimental average Nusselt

number with Literature.

This could be because the inline array follows the square web pattern for radial jets which the correlation predicts more than the matrixes of 45 jets. The differences in the values between the inline array and correlation could be attributed to confinement and cross flow effect which degenerates the strength of the jets depending on the level of its impact at a specific Reynolds number and also on experimental errors.

CONCLUSION This investigation has considered inline array of jets for

impingement heat transfer. The results show that uniform distribution of temperature and Nusselt number in the annulus of a concentric tube could be achieved when the jets are uniformly spaced with equal pitch. The homogeneity, though evident, but not so observable could possibly be due to errors in experimental measurements

In [19], it was recommended that for there to be good interaction between an array of impinging jets on a surface, the pitch of the jets 𝑆 should be equally spaced and should be varied between 1.0− 1.5𝐻 from the impingement surface. It was also recommended that for optimum heat transfer for array of round jets, 𝐷 = 0.2𝐻 and 𝑆 = 1.4𝐻, where 𝐻 is the fixed parameter in this case of asymmetry. The distance or height of the holes drilled in the cooling tube from the impingement surface should be within the potential core of a turbulent jet which is around 6 to 8 jet diameters. Opposite rows of jets could be spaced on the tube at 90 degrees as this has shown to give good effect. The exhaust could be designed on the impingement surface and the spent fluid could be made to exit on both sides of the concentric tube if cross flow effects should be minimised. The ratio of the area of the exhaust to the flow area in the annulus should not be less than 1 for good results. Martin [19] recommended an area ratio of 3. Since increase in Reynolds number in the turbulent range in the concentric annulus basically enhances mixing, power can be saved by observing at what Reynolds number the mixing is uniform as further increase would not be necessary.

ACKNOWLEDMENT Special thanks to Dr. Roger Penlington of the Mechanical

Engineering department, Northumbria University who was very instrumental as my supervisor during the course of this research.


REFERENCES [1] Bernia, M. et al. (1999), “Numerical study of turbulent heat transfer in confined and unconfined impinging jets”, Journal of Heat and Fluid Flow, 20(1), pp. 1-9. [2] Brignoni, L. A. and Garimella, S. V. (1999), “Effects of nozzle-inlet chamfering on pressure drop and heat transfer in confined air jet impingement”, Journal of Heat and Mass transfer, 43(7), pp. 1133-1139. [3] Chang, S. W. et al. (2006), “Heat transfer of impinging jet-array over convex-dimpled surface”, Journal of Heat and Mass Transfer, 49(17-18), pp. 3045-3059. [4] Chang, S. W. et al. (2009), “Jet-array impingement heat transfer in a concentric annular channel with rotating inner cylinder”, Journal of Heat and Mass Transfer, 52(5-6), pp. 1254-1267. [5] Chattopadhyay, H. (2004), “Numerical investigations of heat transfer from impinging annular jet”, Intern. Journal of Heat and Mass Transfer, 47(14-16), pp. 3197-3201. [6] Florscheutz, L. W. (1981), “Streamwise flow and heat transfer distributions for jet array impingement with cross flow”, Journal of Heat Transfer, 102, pp. 337–342. [7] Florschuetz, L. W. et al. (1984), “Heat transfer characteristics for jet array impingement with initial crossflow”, ASME Journal of Engineering for Power, 106 (1), pp. 34-41. [8] Garimella, S. V. and Nenaydykh, B. (1996), “Nozzle-Geometry Effects in Liquid Jet Impingement Heat Transfer”, Journal of Heat and Mass Transfer, 39(14), pp. 2915-2924. [9] Goodro, M. (2008), “Effects of hole spacing on spatially-resolved jet array impingement heat transfer”, Journal of Heat and Mass transfer, 51(25-26), pp. 6243-6253. [10] Haiyong, L. et al. (2008), “The Effect of Jet Array Arrangement on the Flow Characteristics of the Outlet Hole in Short Confined Channels”, Heat Transfer-Asian Research, 37(1), pp. 20-27. [11] Hollworth, B. R. and Dagan, L. (1980), “Array of Impinging Jets with Spent Fluid Removal through Vent Holes on Target Surface – Part 1: Average Heat Transfer”, ASME Journal of Engineering for Power, 102, pp. 994-999. [12] Huang, C. L. et al. (1997), “Local Thermal characteristics of a Confined Round Jet Impinging onto a Heated Disk”, Electronic Packaging Technology Conference, pp. 108-114. [13] Huber, A. M. and Viskanta, R. (1994), “Impingement heat transfer with a single rosette nozzle”, Experimental Thermal and Fluid Science, 9(3), pp. 320-329. [14] Huber, A. M. and Viskanta, R. (1994), “Convective heat transfer to a confined array of air jets with spent air exits”, Journal of Heat Transfer, 116(3), pp. 570-576. [15] Incropera, F. P. and DeWitt, D. P. (1990), Fundamentals of Heat and Mass Transfer 3rd edn. Canada: John Wiley & Sons, Inc. [16] Jambunathan, K. et al. (1992), “Review of heat transfer data for single circular jet impingement”, International Journal of Heat and Fluid Flow, 13(2), pp. 106–15. [17] Koopman, R. N. and Sparrow, E. M. (1976), “Local and Average Transfer Coefficients due to an Impinging Rows of

Jets”, Int. Journal of Heat and Mass Transfer, 19 (6), pp. 673-683. [18] Lee, D. K. and Vafai, K. (1999), “Comparative Analysis of Jet Impingement and Micro channel cooling for High Heat Flux Application”, International Journal of Heat and Mass Transfer, 42, pp. 1555-1568. [19] Martin, H. (1977), “Heat and Mass Transfer between Impinging Gas Jets and Solid Surfaces”, Advances in Heat and Mass Transfer, 13, pp. 1-60. [20] Pan, Y. and Webb, B. W. (1995), “Heat Transfer Characteristics of Arrays of Free-surface Liquid Jets”, ASME Journal of Heat Transfer, 117(4), pp. 878-883. [21] Polat, S. (1993), “Heat and mass transfer in impingement drying”, Drying Technology, 11(6), pp. 1147-1176. [22] Polat, S. and Douglas, W. J. M. (1990), “Heat transfer under multiple slot jets on a permeable moving surface”, AICHE Journal, 36(9), pp. 1370-1378. [23] Royne, A. and Dey, C. J. (2006), “Effect of nozzle geometry on presure drop and heat transfer in submerged jet arrays”, Journal of Heat and Mass Transfer, 49(3-4), pp. 800-804. [24] San, J. Y. and Lai, M. D. (2001), “Optimum jet-to-jet spacing of heat transfer for staggered arrays of impinging air jets”, Journal of Heat and Mass Transfer, 44(21), pp. 3997-4007. [25] Saripalli, K. R. (1983), “Visualization of Multijet impingement flow”, AIAA Journal, 21 (4), pp. 483-484. [26] Sarkar, A. et al. (2004), “Fluid Flow and Heat Transfer in Air Jet Impingement in Food Processing”, Journal of Food Science, 69(4), pp. 113-122. [27] Sharif, M. A. R. and Banerjee, A. (2009), “Numerical analysis of heat transfer due to confined slot-jet impingement on a moving plate”, Applied Thermal Engineering, 29(2-3), pp. 532-540. [28] Slayzak, S. J. et al. (1994), “Effects of Interaction between adjacent free-surface planar jets on local heat transfer from the impingement surface”, International Journal for Heat and Mass Transfer, 37(2), pp. 2689-282. [29] Su, L. M. et al. (2003), “Impingement Heat Transfer of Reciprocating Jet Array”, JSME International Journal, 46(3), pp. 1-17. [30] Wang, T. et al. (2005), “Flow and heat transfer of confined impinging jets cooling using a 3-D transient liquid crystal scheme”, 48(23-24), pp. 4887-4903. [31] Webb, B. and Ma, C. F. (1995), “Single phase liquid jet impingement heat transfer”, Advances in Heat transfer, 26, pp. 105-217. [32] Weigand, B. and Spring, S. (2009), “Multiple Jet Impingement – A Review”, International Symposium on Heat Transfer in Gas Turbine Systems, 9(14), pp. 1-3.




FINITE ELEMENT ANALYSIS OF A HIGH PERFORMANCE HEAT SINK WITH COMPLEX GEOMETRY IN POWER ELECTRONIC BUILDING BLOCKS

Piero Caballero, Sam Yang, Alejandro Rivera, and J. C. Ordonez

Department of Mechanical Engineering and Center for Advanced Power Systems Florida State University Tallahassee, FL, USA

ABSTRACT

This paper presents a finite element analysis employed to find the overall convective heat transfer coefficient of a high performance heat sink in a Power Electronic Building Block. Due to the complex geometry of the heat sink, currently available empirical correlations for the convective heat transfer coefficient are not applicable. As a result, COMSOL Multiphysics software and parameter estimation by least squares approach are used as the main simulation and analysis tools to find the coefficient, in addition to the experimental temperature measurements of the heat sink under a specific heat load. Simulation was performed with a range of convective heat transfer coefficients until the average temperature of the simulated heat sink matched to that of the experimental one. In this study, the overall convective heat transfer coefficient of the heat sink is determined to be 19.62 W·m-2·K-1.

INTRODUCTION Power electronics has become one of the most prominent

research areas since its first appearance, with a wide variety of solid-state electronic applications for the electric power control and conversion. Power electronics aims to effectively transmit the electrical power, and its forefront research areas involve the study and optimization of electronic systems to increase the conversion efficiency and the quality of the transmitted power. Consequently, the development of such systems raises issues on thermal management since most of the power loss is directly proportional to heat generated by the system during the conversion. Power losses dissipated as heat are often negligible compared to the total power converted; however, these losses must be taken into account when dealing with large systems such as an all-electric ship. Furthermore, most power electronic devices are confined to enclosures, which can often be a challenge because improper thermal management in such systems may result in thermal runaway of electronic devices, leading to a system failure. Therefore it is essential to develop an effective method to manage the heat generated, in order to prevent potential system faults and increase system reliability by satisfying all cooling requirements.

For any cooling system, it is essential to determine the amount of heat transferred from the heat sink to the cooling fluid (i.e. air, water, etc.), which heavily depends on the heat transfer area and the overall convective heat transfer coefficient, hc. As a result, this work attempts to use COMSOL Multiphysics 4.3b (COMSOL) to characterize the thermal behavior of a passive high performance heat sink in a power electronic building block (PEBB) by finding its hc.

A PEBB features several electronic components as illustrated in Figure 1. The primary goal of this PEBB is to convert electric power from AC to DC with three 150 A thyristor modules while controlling three-phase voltage and current, maintaining a favorable operating condition for all electronic devices in it. Thyristors are important devices in AC circuits, as they switch between forward and reverse bias (i.e. light dimmer, power controller, etc.). However, they tend to dissipate a large amount heat, and therefore a heat sink is placed inside the PEBB to manage the heat generated by these thyristor modules to the surroundings as shown in Figure 2.

Figure 1. Diagram of a PEBB with its primary electronic

components [1].


Figure 2. PEBB experimental set-up with thyristors attached to the heat sink.

In an effort to enhance the cooling of the thyristors, a 300

mm extrusion heat sink manufactured by AAVID Thermalloy (Model No. 62905) is integrated into the PEBB, and hence employed in this study. Figure 3 exhibits the 62906’s fin surfaces extended in shapes of tree branches or river basins, which resembles those obtained using constructal theory [2]. The AAVID heat sink is made of 6063-T5 aluminum with a thermal resistance of 0.55 ⁰C·W-1, and Figure 4 shows a three-dimensional heat sink model created for this analysis.

Figure 3. AAVID Thermalloy 62905 heat sink; dimensions are in millimeters.

Figure 4. 3D model of an AAVID Thermalloy 62905 heat

sink.

MATHEMATICAL BACKGROUND The following assumptions are made in effort to simplify

the problem and reduce the overall simulation time:

1. Steady-state 2. No internal heat generation 3. Radiation heat transfer is neglected 4. Constant thermal conductivity and ambient temperature.

Based on these assumptions, the first governing equation in fin analysis is the Laplace equation:

∇2T = 0 (1)

The boundary condition implemented at the surface of the heat sink can expressed as:

−𝐧 · (−k∇T) = h(T− T∞) (2)

which accounts for the convection occurring at the boundary of the heat sink with the ambient based on Newton’s law of cooling and Fourier’s law.

The mathematical expression for the boundary heat flux from the three thyristor modules is defined as:

−𝐧 · (−k∇T) = q′′ (3) where q′′ is the total power dissipated by the three thyristor modules at the boundary per unit surface area. The boundary heat flux is only applied on the surfaces where thyristor modules are located.

COMSOL MODELING AND SIMULATION The complex geometry of the heat sink makes it difficult to

calculate hc based on presently available empirical correlations. Therefore, the heat sink is modeled in COMSOL and a finite element analysis is performed with various values for hc to obtain the heat sink temperature under the same heat load as the experiment.

In the experiment, temperatures at three different locations on the heat sink, as indicated in Figure 5 with red dots, are recorded using a data acquisition system composed of LabView 8.2, NI PXI 1010 – Chassis, PXI-6251 – Data Acquisition, and high precision thermistors of type 44004RC, standard type Bead 1, with a maximum diameter of 2.4 mm. For calibration purposes, the thermistors are immersed in a constant temperature fluid, and sixty-four temperature measurements are made starting at 20 ⁰C and up to 80 ⁰C by increments of 10 ⁰C. The largest standard deviation of these measurements was 0.0006 ⁰C, thus the bias limit is considered to be ±0.0012 ⁰C for all the thermistors. Consequently, the uncertainty in all temperature measurement is determined to be ±0.2 ⁰C [3]. In order to ensure good thermal contact between the thermistors and the heat sink, each thermistor is thermally insulated and


attached to the heat sink as shown in Figure 6. The average steady-state temperature (Tavg) of these three points is measured to be 312 ± 0.2 K.

Figure 5. Top view of an AAVID heat sink where red dots

indicate the positions of thermistors.

Figure 6. Thermal Insulation around the thermistor of type 44004RC.

In order to replicate the thermal behavior of an actual

AAVID heat sink, material properties of 6063-T5 aluminum are integrated into the COMSOL material database. Additionally, the boundary heat flux and locations of thyristor modules are excerpted from the previous PEBB experimental setup [1, 3], where the total heat dissipated by three thyristor modules is calculated to be 182 W. The initial temperature of both the heat sink and the immediate surrounding is assumed to be at 296 K. Based on the coordinates from the experimental setup, temperature probes are virtually installed in COMSOL at exactly the same locations as the thermistors in the experiment (refer to Figure 5). Then the steady-state Tavg of these points in COMSOL is calculated and compared to the experimental measurement.

The model is simulated in COMSOL with a set of hc using the parameter sweep. The first set of hc is assigned to range from 0 to 30 W·m-2·K-1, with an increment of 1.0 W·m-2·K-1 in between. Then a smaller range of heat transfer coefficients is selected based on the accuracy of the simulated Tavg with respect to the experimental Tavg. As a result, the new range of hc is set to be from 19.55 to 19.65 W·m-2·K-1, with a step size of 0.01 W·m-2·K-1.

In addition to the parameter sweep, mesh convergence is assessed in order to validate the accuracy of the average temperature obtained from the finite element analysis. The

same set of hc is simulated repeatedly with increasing numbers of elements and mesh quality, and as a result, a more accurate hc is determined. In this study, a free tetrahedral mesh is used to compensate the complex geometry of the fins. Table 1 displays mesh convergence results, where R is the mesh refinement number, Nelem, Nedge and Nvert are the total number of elements, edges and vertices in the mesh, respectively, and DOF is the number of degrees of freedom, which is the number of unknowns for the finite element method that needs to be solved for.

Table 1. Mesh convergence results.

R 1 2 3 Nelem 393123 931974 1894061 Nedge 41850 66682 88349 Nvert 913 913 913 DOF 697702 1561850 3042624

Parameter estimation is essential in determining a more

accurate estimation for the target value based on the least squares method. Such estimation involves a squared deviation calculation which is expressed as:

ε = Σ (ΔTavg)2 (4) where ΔTavg is the difference between the simulation and experimental average temperatures of the three points indicated previously. Parameter estimation is employed for each mesh refinement number, R, and the results are shown in Figure 7.

Figure 7. Least squares plot to find an accurate hc with

different mesh refinement number R.

In Figure 7, the absolute minimum is where the desired value for the hc is located, which is evaluated to be 19.62 W·m-

2·K-1. The mesh convergence study shows that hc becomes more accurate from R = 1 to R = 2; however, refining the mesh further does not reduce the error. Thus, R = 2, as shown in Figure 8, is sufficiently refined for this specific geometry.


Figure 8. Illustration of the mesh when R = 2.

The three-dimensional temperature gradient of the heat sink is illustrated in Figure 9.

Figure 9. COMSOL simulation results showing the steady-state temperature gradient of the AAVID 62906

heat sink.

CONCLUSION This work illustrates a finite element analysis performed in

COMSOL in effort to determine the overall convective heat

transfer coefficient of a high performance heat sink with complex geometry, for which currently available correlations for hc cannot be used. Therefore, experimental measurements, parameter sweep, parameter estimation, and mesh convergence study are employed to find an accurate hc for the heat sink, which is determined to be 19.62 W·m-2·K-1. Future work may involve a comparative study between tree branch shaped and flatback shaped heat sinks and an evaluation of the thermal performance of each. Furthermore, the heat sink employed in this study can be modeled with an air flow to assess its thermal behavior under different air velocities.

ACKNOWLEDGEMENT This work was supported in part by the Office of Naval

Research (ONR), the Naval Engineering Education Center (NEEC), and the Multi-physics of Active Systems and Structures (MASS) NSF REU Summer Program.

REFERENCES [1] Ordonez, Juan C.; Rivera, Alejandro; Yang, Sam; Shah, Darshit; Delgado, David; Coleman, Michael; Dilay, Emerson; Vargas, Jose V.C., "Thermal management aspects of all-electric ships," Electric Ship Technologies Symposium (ESTS), 2013 IEEE , vol., no., pp.55,61, 22-24 April 2013. [2] A. Bejan and S. Lorente, The constructal law of design and evolution in nature, Philosophical Transactions of the Royal Society, 365 (2010). [3] E. Dilay, J.V.C. Vargas, J.C. Ordonez, S. Yang, R. Schrattenecker, M. Coleman, T. Chiocchio, J. Chalfant, C. Chryssostomidis, The experimental validation of a transient power electronic building block (PEBB) mathematical model, Applied Thermal Engineering, Volume 60, Issues 1–2, 2 October 2013, Pages 411-422, ISSN 1359-4311.




GAS TURBINE EXHAUST-HEAT-DRIVEN COOLING CYCLE FOR STEAM CONDENSER COOLING AIR APPLICATION

Muhammad M. Mahmood General Electric

Dubai, United Arab Emirates

Drake Viscome General Electric

Greenville, SC, USA Sheldon M. Jeter

Woodruff School of Mechanical Engineering Georgia Institute of Technology

Atlanta, GA, USA

ABSTRACT Combined cycle power plants are much more efficient

compared to simple cycle plants. Combined cycle plants usually consume water for cooling in the steam condenser. In a country like Saudi Arabia the cost of water is high. This leads the utility companies to accept degraded efficiency and install a simple cycle system as it is more economical overall. A combined cycle plant that uses less or no water at all will be a great advantage in this scenario. The idea proposed is to use the waste heat in the exhaust gases of the gas turbine to drive a heat driven refrigeration cycle. The cooled air can then be used for cooling the steam condenser.

This paper presents an integrated open cycle with cooling by humidification cycle and moisture recovery. A mere absorption cycle is not deemed sufficient. The cycle is modeled in EES and ASPEN. A very close agreement is seen. The findings indicate the waste heat in a combined cycle configuration has sufficient potential to achieve water savings in the steam condenser. Although, further work is required for a clearer understanding of how many bottoming cycles can be had whilst being economical.

BACKGROUND OF THE APPLICATION In combined cycle power plants, heat is rejected by the

steam condenser. The steam condenser cooling method of choice is evaporative cooling. Evaporative cooling is the most economical as it is less dependent on the ambient temperature than a dry condenser and also allows lower condenser temperatures for better performance.

Usage of water in combined cycle power plants for evaporative cooling in arid regions can be expensive due to the cost required to procure the water. Usage of desalinated water is expensive, and depleting of precious ground water leads to environmental and social concerns. Hence, any means by which the consumption of water for evaporative cooling in cooling towers is eliminated or reduced is going to be very beneficial.

This would allow nations like Saudi Arabia to build combined cycle plants in arid regions instead of inefficient simple cycle power plants.

DESCRIPTION OF THE CYCLE The cooling cycle that is proposed is not a mere absorption

cycle. It uses cooling by humidification, which is much more effective in reducing the dry bulb temperature. The humidification is followed by moisture recovery so that overall the cycle remains water neutral.

The schematic of the proposed cycle is shown in the figure below:

Figure 1. Cooling Cycle Schematic

The simulation of a cycle very similar to the above is

presented by Hellman and Grossman [1]. An experimental study of the same cycle has also been conducted and available in literature [2], and this is a powerful alternative that combines evaporative cooling and the absorption cycle. The cycle proposed in this paper, however, is different because it employs moisture recovery and therefore does not require an external source of water. The advantages of an open cycle are that low temperature heat sources can be used, and also it eliminates the use of more expensive pressurized heat exchangers.


The cycle utilizes a mixture of Lithium Bromide (LiBr) and water as the refrigerant. LiBr is the chosen absorbent, as it is commonly applied in commercial systems, [1] and a considerable literature is available on cycles based on it. LiBr properties are also available in EES software that are based on the formulation of Patek and Klomfar [3].

The cycle takes in waste heat from the gas turbine exhaust gases leaving the Heat Recovery Steam Generator (HRSG) at state 13. These gases heat the weak solution in the generator which enters at state 143. The cooled exhaust leaves at state 14 and the resulting strong solution and steam leaves at states 140 and 102 respectively. The steam is then condensed by use of ambient air in the condenser to state 103 at slightly above ambient temperature. The condensed water then mixes adiabatically with ambient air that enters the evaporator at state 110. The cooled air that has its dry bulb temperature very close to the wet bulb temperature, then transfers its cooling to an ambient air stream at state 120 in an air-to-air heat exchanger. The cooled air at state 121 then cools the steam in the steam condenser. Finally, the hot and moist air at state 113 is heated to a temperature slightly above the ambient before it enters the absorber for dehumidification. The absorber is operating at an isothermal temperature slightly above the ambient, so it is able to reject heat to the ambient air. The strong solution at state 140 makes its way to the absorber by first preheating the weak solution at state 143, which is leaving the absorber. The strong solution also then is cooled to the absorber temperature by the hot moist air stream.

All of the state points mentioned above, except for exhaust gas temperature entering and exiting the generator, can have a range of temperatures. The final set points are to be based eventually on a tradeoff between economics and Coefficient of Performance (COP). For instance, operating the absorber at temperatures approaching the ambient temperature increases the COP but leads to a very high mass flow of air required for cooling. This leads to higher parasitic power losses in terms of fan power and a much larger absorber. Hence, the optimum set points are to be based on a detailed economic evaluation. In this paper the state points have been somewhat arbitrarily fixed to yield a realistic COP and at air mass flow rates and heat exchanger overall heat transfer coefficient (UA) values that are not unrealistically high. Table 1 shows the typical temperatures that can be expected without a thorough optimization based on an economic evaluation. With values similar to these, a COP of 0.75-0.85 is achieved. The exhaust gas entering temperature is based on the simulation of a combined cycle model in EES which will be explained in details in the following sections.

EES MODEL The cooling cycle is simulated by programming the energy

and mass balances for each component of the cycle in EES. There are several unknowns which will render the energy balance and mass balance equations insolvable. Therefore, assumptions have to be made for several parameters. Also, there are quantities that are known and easily determined which

will, together with the assumptions, lead to solving the entire system of equations.

The exhaust gas entering temperature is a known quantity, and the temperature of the exhaust gases leaving can also be fixed to a certain temperature which will prevent condensation from occurring in the stack. The mass flow rate of the exhaust gases is also known from the simulation of the combined cycle. These values allow easy computation of the heat supplied to the cooling cycle. The heat in the absorber is to be rejected to the ambient. As the absorber is operated isothermally the temperature chosen for the solution and air streams entering the absorber is assumed 3K above the ambient.

The weak and strong solution concentrations on a mass basis are key unknowns. The concentration is a function of the vapor pressure and temperature of the solution. The weak solution concentration is determined by knowing the absorber temperature and by setting the solution vapor pressure equal to the partial pressure of water vapor in the moist air entering the absorber. Similarly the strong solution concentration is found at the absorber temperature, which is known, and at a vapor pressure which is equal to the partial pressure of water vapor in the dehumidified air leaving the absorber. This partial pressure of water vapor is the same as the partial pressure of water vapor in the ambient air as the humidity ratio of the air leaving the absorber is equal to the ambient. The partial pressure of water vapor is found using the equation below where W is the humidity ratio, 𝑃𝑤 is the partial pressure of water vapor and 𝑃 is the total pressure [4]:

𝑊 = 0.62198𝑃𝑊

𝑃 − 𝑃𝑊 (1)

Table 1. Typical Temperatures Ambient Temperature = 320.9 K Ambient Pressure = 96.42 KPa Relative Humidity = 14.19%

T143 = 345 K T140 = 370 K

T102 = 370 K T103 = 322.9 K

T130 = 320.9 K T131 = 323.9 K

T100 = 320.9 K T101 = 297.4 K

T110 = 320.9 K T111 = 299.4 K

T112 = 320.6 K T120 = 320.9 K

T121 = 319.2 K T122 = 329.8 K

TSteam = 330.8 K T112b = 323.9 K

T113 = 323.9 K T140b = 344.9 K

T141 = 323.9 K T142 = 323.9 K The resulting vapor pressures of the solutions, in terms of

the relative magnitudes, are in line with the typical guidelines in the literature [5]. The highest vapor pressure is achieved after


heat is supplied in the generator. The vapor pressure in the absorber is the lowest and at an intermediate level at the generator entry.

In addition to determining the above mentioned state points and mass flow rates, there are additional assumptions to be made. It is assumed that the water leaving the refrigerant (water) condenser is sub-cooled to 2K above ambient. The cooled moist air is assumed to be at a dry bulb temperature 2K above the wet bulb temperature of the ambient air. With the above mentioned assumptions and known values the following equations for each component are used to simulate the entire cooling cycle:

Generator Energy balance:

��𝑖𝑛 = ��𝑒𝑥ℎ𝑎𝑢𝑠𝑡(ℎ13 − ℎ14) (2)

��𝑖𝑛 = ��140ℎ140 + ��102ℎ102 − ��143ℎ143 (3)

Mass balance:

��𝐿𝑖𝐵𝑟,143 = ��143𝑥143 (4)

Where x143 is the concentration of LiBr on mass basis.

��𝐻2𝑂,143 = ��143 − ��𝐿𝑖𝐵𝑟,143 (5)

��𝐿𝑖𝐵𝑟,140 = ��140𝑥140 (6)

��𝐻2𝑂,140 = ��140 − ��𝐿𝑖𝐵𝑟,140 (7)

��102 = ��𝐻2𝑂,143 − ��𝐻2𝑂,140 (8)

Condenser The mass flow rate in the condenser can be determined by assuming a pinch temperature. An energy balance can then be used to determine the exit state of the air. The general energy balance is:

m102h102 + mairh130 − mairh_131− m102h103 = 0 (9)

Evaporator In the evaporator the assumed value is the dry bulb temperature of the moist air leaving. It is assumed to be just 2K above the wet bulb temperature of the entering air. The mass flow rate of the air and other unknowns are found using energy and mass balances as follows: Energy balance:

��110 =��110ℎ111 + ��101ℎ101 − ��100ℎ100

ℎ110 (10)

Mass balance:

��110𝑊111 = (��110𝑊110) + (��100 − ��101) (11)

Absorber As described earlier the main states for the absorber streams are known. The major unknowns to solve for are the cooling air mass flow required and the heat rejected in the absorber. Energy balance:

��112𝑏ℎ112𝑏 − ��113ℎ113 + ��141ℎ141 − ��142ℎ142− ��𝑟𝑒𝑗𝑒𝑐𝑡𝑒𝑑 = 0

(12)

��133ℎ133 − ��132ℎ132 = ��𝑟𝑒𝑗𝑒𝑐𝑡𝑒𝑑 (13)

Mass balance:

��𝐻2𝑂,142 = ��𝐻2𝑂,141 + ��112(𝑊112 −𝑊113) (14)

The remaining components are the heat exchangers. For

these, simple energy balances are used to determine an unknown thermodynamic state.

1ST AND 2ND LAW CHECKS OF THE MODEL To ensure that the model is not giving rise to cases which

are thermodynamically impossible, 1st and 2nd law checks are done. These are performed on each component and the entire control volume that encloses the cooling cycle.

The second law check requires the calculation of entropy generation for each component in the cycle. The entropy generation must be positive. An important advantage of doing second law analysis is that it gives insight into the location of maximum exergy destruction (or entropy generation). The maximum exergy destruction is found in the generator, where a relatively high temperature source is utilized to generate steam


at a lower temperature. This is an important finding, which sheds light towards possible areas of improvements, and these are discussed in the conclusion section of the paper.

ASPEN MODEL ASPEN is a commercial package most commonly used to

simulate chemical processing plants particularly for the oil and gas and the petrochemical sector. Modeling the cooling cycle in ASPEN and obtaining a close agreement with EES will validate the EES analysis.

The most important step in ASPEN is to choose an appropriate property method to simulate the LiBr solution and extract its thermodynamic properties. The most accurate method in literature [6] [7] to model a LiBr solution is to model it as an electrolyte. There are dedicated property methods in ASPEN to work with electrolytes and the property method selected is ENRTL-RK. As will be described in the next section, this yields a very close agreement with the EES model. The electrolyte wizard in ASPEN is used to model the only required reaction, which is the association/disassociation of LiBr.

ASPEN uses basic blocks in combination with input and output streams to allow construction of a model. The model for each of the cooling cycle components will be described in this section. The generator model is shown below:

Figure 2. Generator Model

The generator is not modeled using the multi-stream heat

exchanger block but is modeled using two heater blocks as it is more convenient for this purpose and is an acceptable method [8]. The first heater simply cools the exhaust gases to a given state at 14 whilst releasing heat. This heat then heats up the weak solution at 143 and produces a mixture of strong solution and steam. This mixture is then separated using the separator block in ASPEN.

The following figure shows the condenser model:

Figure 3. Condenser Model

The condenser is modeled using the two-stream heat exchanger. The shortcut method in ASPEN is used to simulate this heat exchanger. With the same stream conditions as in EES the model yields a heat exchanger UA value very close to that calculated in EES.

The next figure shows the model of the evaporator and the air-to-air heat exchanger that follows the evaporator:

Figure 4. Evaporator & Air-to-Air Heat Exchanger Model

The evaporator is modeled in ASPEN using the adiabatic

mixer block. The ambient air stream and the condensed water are mixed here adiabatically with results closely matching those in EES.

Finally the absorber is modeled. The absorber is very similar to the generator and it is shown in the figure below:

Figure 5. Absorber Model

The absorber is based on heaters and separators just as the

generator is. The heater (streams going from right to left) on the top mixes the strong solution with moist air at ambient pressure. The mixture is then separated into the weak solution and air to be exhausted to the ambient. The weak solution is then throttled to the pressure that it needs to get back to the state at which it enters the generator. This process releases some heat and water vapor which is separated from the stream in an additional separator.

In ASPEN the cycle needs to be “cut” at a certain point to allow input of values to completely define the model. This cycle is therefore considered to start with the stream that defines the weak solution entering the generator, and it terminates at the weak solution leaving the absorber. If these states match within a small tolerance, then it can be assumed that the cycle is performing as expected.


EES AND ASPEN COMPARISON One of the key aims of this paper is to compare the results

of the cooling cycle simulation in EES and ASPEN. A close agreement in the two software packages would add credibility to the modeling approach. For each state point that is present in Figure 1 a comparison is made and the results displayed in the tables that follow.

Table 2. Comparison of EES and ASPEN: Solution

As can be seen in the table above, in most cases the

agreement between EES and ASPEN is very close. In the cases where the difference is really significant, this is due to the fact that the absorption process in ASPEN allows the absorption of air into the strong solution. The absorption of air into the LiBr solution is not modeled in EES, and hence in those state points the difference in vapor pressure and temperature is more significant. The table that follows shows the comparison for the steam/water streams in the cycle:

Table 3. Comparison of EES and ASPEN: H2O Streams

Table 4 shows the comparison of EES and ASPEN data for the moist air streams in the cycle. It can be seen from both of these that the difference between the two models is very small. Hence, it is concluded that both models are describing the same cycle thermodynamically.

Table 4. Comparison of EES and ASPEN: Moist Air Streams

MODEL OF THE COMBINED CYCLE PLANT This section presents the combined cycle model that was

created in EES. The purpose of creating this model is to determine the temperature of the exhaust gases that enter the generator. The combined cycle that is modeled consists of 2 gas turbines, 1 steam turbine with reheat and a single pressure Heat Recovery Steam Generator (HRSG).

The gas turbine model has several simplifying assumptions. All of the compressed air is assumed to be available for combustion, combustors have negligible pressure drop, and compressor and turbine efficiency is 86%. Most of the specifications required to complete the gas turbine model, such as the turbine inlet temperature, are taken for the GE 7FA combustion turbine from publicly available manuals [9] [10]. For the HRSG the assumption includes pinch temperature and approach temperature, as illustrated in Figure 6. The boiler pumps efficiency is assumed at 85%. The heat rate diagram in Figure 6 summarizes the HRSG.

State # Parameters EES Model ASPEN Model % Error

T [K] 345 345 0.00%P [KPa] 8.4 8.6 2.33%

T [K] 370 370 0.00%P [KPa] 14.7 15.4 4.55%

T [K] 344.9 344.9 0.00%P [KPa] 4.7 5 6.00%

T [K] 323.9 323.9 0.00%P [KPa] 1.568 1.7 7.76%

T [K] 323.9 323.9 0.00%P [KPa] 2.964 3.9 24.00%

T [K] 345 345 0.00%P [KPa] 8.398 7.6 9.50%

Solution

143

140

140b

141

142

142b

Comparison of EES & ASPEN


T [K] 370 370 0.00%P [KPa] 14.7 15.4 4.55%

T [K] 322.9 322.9 0.00%P [KPa] 14.7 15.4 4.55%

Steam/Water


102

103


T [K] 457.9 457.9 0.00%P [KPa] 96.42 96.42 0.00%

T [K] 400 400 0.00%P [KPa] 96.42 96.42 0.00%

T [K] 320.9 320.9 0.00%P [KPa] 96.42 96.42 0.00%

T [K] 299.4 298.85 0.18%P [KPa] 96.42 96.42 0.00%

T [K] 320.6 319.9 0.22%P [KPa] 96.42 96.42 0.00%

T [K] 323.9 323.15 0.23%P [KPa] 96.42 96.42 0.00%

T [K] 323.9 323.9 0.00%P [KPa] 96.42 96.42 0.00%

T [K] 320.9 320.9 0.00%P [KPa] 96.42 96.42 0.00%

T [K] 319.2 319.2 0.00%P [KPa] 96.42 96.42 0.00%

T [K] 320.9 320.9 0.00%P [KPa] 96.42 96.42 0.00%

T [K] 323.9 323.35 0.17%P [KPa] 96.42 96.42 0.00%

T [K] 320.9 320.9 0.00%P [KPa] 96.42 96.42 0.00%

T [K] 323.8 323.25 0.17%P [KPa] 96.42 96.42 0.00%

130

131

132

133

121

111

112

112b

113

120

13

14

110

Air



Figure 6. HRSG Heat Rate Diagram

A detailed combustion equilibrium model is also created.

The fuel for the gas turbine is assumed to be 100% methane. Also, NOx formation in the combustion process is assumed to be negligible. The schematic for the gas turbine model is shown below:

Figure 7. 7FA Gas Turbine Schematic

The schematic for the HRSG and the steam turbine is

shown in the schematic below:

Figure 8. HRSG & Steam Turbine Schematic

CONCLUSIONS This cycle has been developed and given preference over

other similar configurations as it can be used to recover low

grade heat from the exhaust and also it is water neutral overall. The use of cooling air by evaporation of water is an effective means to achieve significant cooling specially in hot and dry areas of low humidity. The analysis has revealed that the ambient temperature, pressure and humidity of the air are key factors and will dictate the concentrations for the LiBr solutions. This will in turn affect the entire cycle performance.

Without thorough attempts at optimizing the cycle with respect to maximizing the cooling or minimizing the cost a COP of 0.75-0.85 is achieved. This COP is not sufficient to provide enough cooling for condensing the steam. At best, if maximum heat is extracted from the exhaust gases, the cooling achieved is about 30% of the required cooling for the steam condenser.

The most important conclusion to be drawn is that a single stage of this cooling cycle is not sufficient. The highest exergy destruction is seen in the generator. This is primarily due to utilizing high temperature exhaust to produce steam at lower temperature. A low temperature is favorable as it leads to more water production for cooling.

An alternate is to have multiple stages of the cooling cycle. The generator can operate at a high temperature and produce steam at higher temperatures that are approaching the exhaust. This high temperature steam has enough thermal exergy content to then be utilized as a heat input stream for another generator. In the current proposal the thermal exergy of the steam produced is rejected to the ambient in the condenser. With multiple stages a COP of around 4 can be achieved which will provide sufficient cooling for the condenser and hence eliminate the need of any water.

REFERENCES [1] Hellmann, H.M., and Grossman, G., 1995, ‘‘Simulation

and Analysis of an Open-cycle Dehumidifier-Evaporator Regenerator (DER) Absorption Chiller for Low-grade Heat Utilization,’’ Int. J. Refrigeration., 18, pp. 177–189.

[2] Gommed, K., Grossman, G. and Ziegler, F., 2004, ‘‘Experimental Investigation of a LICL-Water Open Absorption System for Cooling and Dehumidification,’’ ASME J. Solar Energy Engineering, 126, pp. 710–715.

[3] Patek, J., and Klomfar, J., 2006, “A Computationally Effective Formulation of the Thermodynamic Properties of LiBr-H2O Solution from 273 to 500 K Over Full Composition Range,” Int. J. Refrigeration., 29, pp. 566-578.

[4] ASHRAE Handbook – Fundamentals SI, Chapter 6 Psychrometrics, 2005, American Society of Heating, Refrigeration and Air-Conditioning Engineers.

[5] ASHRAE Handbook – Fundamentals SI, Chapter 32 Sorbents and Desiccants, 2009, American Society of Heating, Refrigeration and Air-Conditioning Engineers.

[6] Somers, C., 2009, “Simulation of Absorption Cycles for Integration into Refining Processes,” MSc thesis, University of Maryland, College Park, MD.


[7] Aspen Plus Manual, 1999, “Modeling Processes with Electrolytes,” AspenTech.

[8] Scheflan, R., 2011, “Teach Yourself the Basics of Aspen Plus,” John Wiley & Sons.

[9] Jacobs, J.A., and Schneider, M., “GER-3430G Cogeneration Application Considerations,” General Electric Reference, http://site.ge.energy.com/prod_serv/products/tech_docs/en/all_gers.htm

[10] Chase, D.L., and Kehoe, P.T., “GER-3574G GE Combined-Cycle Product Line and Performance,” GE Ref., http://site.ge.energy.com/prod_serv/products/tech_docs/en/all_gers.htm




SERPENTINE PARTICLE-FLOW HEAT EXCHANGER WITH WORKING FLUID, FOR SOLAR THERMAL POWER GENERATION

Matthew Golob, Dennis Sadowski, Sheldon Jeter The George W. Woodruff School of Mechanical Engineering

Georgia Institute of Technology Atlanta, Georgia, USA

ABSTRACT This experimental study is conducted as part of a DOE-

funded SunShot project titled “High Temperature Falling Particle Receiver”. A 300kW-th Concentrated Solar Power Tower is currently in development, in which silica sand acts as the heated medium. Sand is passed through a solar receiver, where it absorbs concentrated solar irradiation, and subsequently flows through a heat exchanger with the purpose of transferring thermal energy to a working fluid. The working fluid will, in turn, flow through a power-generation device, such as a turbine system.

An experiment with a near-scale Serpentine Particle-Flow Heat Exchanger system was constructed with the primary purpose of measuring the effective heat transfer coefficient. This was to quantify a heat transfer coefficient between mid-grain construction sand, and a hotter working fluid of air or water. To accomplish this, a closed loop circulation of sand was required to ensure a steady state particle flow condition. The sand was passed though the heat exchanger in mass flow state. An Olds Elevator lifted the sand, allowing the sand to recirculate back into the top of the heat exchanger. The overall effective heat exchange coefficient for the serpentine finned tube exchanger came out to 200-275 W/m2-K at particle flow rates under 0.9 mm/s with a sand side coefficient of 12-17 W/m2-K. These tests have prompted further testing with a revised medium scale finned heat exchanger design with improved flow delivery.

1. INTRODUCTION The idea of central receiver systems based on gas cycles

has gained considerable interest during the past two decades, due to its ability of achieving very high temperatures through its intense heat flux concentration on a relatively small area. Various gas cycle concepts have been proposed and tested, and most involve the direct heating of compressed air or other gas [1-3]. However, one of the major challenges of these systems is the successful incorporation of thermal energy storage, since the effectiveness of using thermal energy storage with air or gas is relatively poor.

There have been a number of thermal energy storage solutions proposed over the past three decades. One of the most widely accepted thermal energy storage solutions is the use of molten salts [4,5]. Currently, the use of molten salts for thermal energy storage is limited to temperatures generally less than

600°C due to technical restrictions in piping and salt phase stability. Another solution is the use of solid blocks to store energy during the day. This concept has been demonstrated with concrete blocks [6,7], but the temperatures are generally limited to less than 500°C due to concrete properties, making this model unsuitable for high-temperature applications. Furthermore, since solid blocks store sensible heat, their temperature profile during the discharging process causes a gradual decline in cycle efficiency. Yet another solution is to use sand as a storage medium [8]. This concept was developed to work in conjunction with an air receiver. The sand is heated in an air-sand heat exchanger to a very high temperature. The sand then flows to a hot storage tank, and then to a fluidized bed cooler, where its heat is used to generate steam that feeds a steam power cycle. The colder sand returns either to the air-sand heat exchanger or is stored in a cold storage tank. This technology resolves the temperature limit issues faced in the solid block concept. However, the main issue of the unfavorable temperature profile during discharging still persists.

Figure 1.1. Tower Concept Configuration

To overcome the thermal energy storage issues described

above, a promising solution is being developed by researchers at King Saud University and the Georgia Institute of Technology. It involves the use of sand or other fine granular materials as the primary thermal medium. Unlike the existing


concepts that only utilize sand for thermal energy storage, the new system allows sand to be directly heated by the incoming sunlight, as Figure 1.1 shows. To reclaim the energy, a portion of the heated sand is flowed through a heat exchanger to transmit thermal energy with compressed air. The now hot air leaving the sand-air heat exchanger is then fed to a suitable gas cycle to generate power. The remainder of the sand is stored in a well-insulated bin where it is kept for later use. When the sand exits the heat exchanger, it has lost much of its energy to the air. The cooled sand is then recirculated to the top of the tower using bucket conveyors or a similar high temperature lift mechanism. During the nighttime, stored hot sand is drawn into the heat exchanger, and the sand leaving the heat exchanger is then diverted to a cold bin where it resides until the solar field comes back in operation the next day. One of the main advantages of this concept is that the hot sand is stored upstream of the heat exchanger. By doing so, the temperature of sand at the beginning of the process of heat exchange will remain high increasing gas cycle efficiency.

The new system, called the high temperature solar gas turbine system, builds on the experience and the outcomes of the solid particle receiver project that was introduced and tested at the National Solar Thermal Test Facility in Sandia National Laboratories [9,10]. The primary expansions on the concept are an improved receiver design and the incorporation of a thermal energy storage unit. To demonstrate the merits of the new concept, a pilot-scale, 300 kW (thermal) central receiver plant will be built on the campus of King Saud University, in Riyadh, Saudi Arabia, Fig. 1.2. Construction is already underway and is expected to be completed in the last quarter of 2013.

Figure 1.2. General Layout of Field and Tower The current study focuses on a single component of the

system, namely the sand-air coiled-finned tube heat exchanger. The aim of this study is to understand the heat transfer characteristics of the heat exchanger such that it can be further designed and refined properly. A number of studies on the heat

transfer between granular material and a flat plate have been carried out, e.g. [11,12]. Studies on the heat transfer between granular material and tube banks were also carried out, two of them are of particular interest [13,14]. The two studies considered different tube arrangements and, therefore, led to significantly different values for the effective heat transfer coefficient of the sand. In reference [13], polypropylene particles ranging in size from 350 to 710 µm were allowed to flow past a single row and two rows of tubes. The local heat transfer coefficient around the tubes of a two-row arrangement was reported to range from approximately 25 to 120 W/m2-K for velocities ranging from 0.4 to 6.7 mm/s. In reference [14], a more extensive tube bank containing 68 tubes was used. Four particle types were studied: ash (average particle size: 475 µm), sand (average particle size: 203 µm), sand (average particle size: 637 µm), and corundum (average particle size: 195 µm). The study reported that the effective average heat transfer coefficient for staggered tube bank ranged from approximately 90 to 220 W/m2-K at various mass fluxes. Of the materials tested, corundum generally showed the highest effective average heat transfer coefficient.

Figure 1.3. Serpentine Particle-Flow Heat Exchanger or Coiled Finned Tube Heat Exchanger

The present study considers bulk flow of sand through a

full scale Serpentine Particle-Flow Heat Exchanger consisting of coiled finned tubes as shown in Fig. 1.3. Several types of sands, that varied in both purity and grain size, were examined to find the most suitable candidate. The experimental setup is described in the next section.

2. SYSTEM PURPOSE AND EXPERIMENTAL SETUP The test system was set up to effectively measure heat

transfer characteristics in the heat exchanger between the sand and working medium of either air or water. To accomplish this, a closed loop circulation of the sand was required to ensure a steady state particle flow condition. For the overall layout, a sketch and photograph of the experimental apparatus is exhibited in Fig. 2.1.


Figure 2.1. Sketch and Photograph of Experimental Apparatus

Sand is passed though the heat exchanger in mass flow

from a hopper located above and into a funnel below the heat exchanger. The sand exits the funnel onto a horizontal conveyor belt suspended on a scale that measures particle mass flow. The conveyor then pours sand into the base of a vertical lift Olds Elevator which raises the sand to a height above the hopper. The sand then exits from the elevator top onto a shallow sloped feed ramp, which with the assistance of a vibrator, pours the sand into the hopper. This closes the loop on the sand side. The hopper, heat exchange box, and discharge funnel are wrapped in insulation to reach a near adiabatic state for that portion of the cycle.

Figure 2.2. Schematic of Air Module System A schematic of the Air Module is displayed in Fig. 2.2.

The primary purpose of the air module is to measure the effective heat transfer coefficient between sand and air to more closely replicate Riyadh Techno Valley project testing conditions. To accomplish this, a once through stream of air is employed to complete the accompanying side of the heat exchange system with the sand recirculation loop. Ambient air is taken into a roots blower through a 4 inch pipe with a 2 inch orifice plate flow meter installed to measure the volumetric flow rate of the air. In the blower the air passes through a filter

and is pressurized to 6 psig with output attached to a 4 inch hose pipe with a temperature tap at the end. The pressurized air is then pushed through a duct air heater bringing the air temperature up to ~200°F. Exiting the air heater, pressure and temperature taps capture the air conditions passing through a 4 inch diameter hose line and entering the sand/air heat exchanger. The air then flows in general counter flow through the coiled finned tubes where it gets cooled by the sand. At the coiled heat exchanger discharged air re-enters the 4 inch hose line where it passes temperature and pressure taps to collect the air discharge conditions. The air is finally ejected to ambient through a diffuser.

The primary purpose of the Water Loop module is to measure a more precise effective heat transfer coefficient in the heat exchanger between sand and water. To accomplish this a closed water loop is used in conjunction with the sand recirculation system. Water is pumped through the module at 25 GPM through a 2 inch pipe. From the pump water passes a pressure tap and flow adjustment valve (cutoff) before entering the water heater. The water passes though the 40 kW circulation heater discharging water at ~180°F. An emergency temperature sensor in the heater exit and pressure relief valve downstream are present incase proper conditions in the water flow are not maintained. Pressure and temperature taps capture the water flow conditions passing through 2 inch diameter pipe before entering the sand/air heat exchanger. The 180°F water then flows in general counter flow through the coiled finned tubes where it gets cooled by the sand. The coiled heat exchanger then discharges the water into a 2 inch hose line where it passes temperature and pressure taps to collect the discharge side conditions of the water. The water then proceeds through to a volumetric tank which maintains the overhead of water for the loop. The water then passed through a water flow meter, expansion tank, and flow control valve to regulate the flow through the loop. A schematic of the Water Loop Module is displayed in Fig. 2.3. It can be observed that this system largely differs from the Air Module by recirculating virtually all working fluid.

Figure 2.3. Schematic of Water Loop Module


3. METHOD AND PARAMETERS For the sand system with air module, test runs involved

setting the blower and air heaters to a particular value matched with a fixed sand flow rate to let the system reach a steady operation over a period of approximately 9-10 hours a day. Three types of sand of differing Sauter Mean Diameter (SMD) grain size were to be tested. The largest grain size, construction sand, was run and the mid grain size sand was also completed. Testing involving the finest grain sized sand was deemed to be too dusty to operate in a continued state. The known characteristics of the three sands, course construction sand with SDM of 355 µm, (wide size range, dusty), mid-grain of SDM of 301 µm (homogenous size, very low dust), and ultra-fine grain sand with SDM of 103 µm (extremely dusty) were compared. From the air-sand results, the mid-grain size was selected as optimal due to its low dust content and heat exchange performance. Testing was only resumed with the mid-grain sized sand for the water-sand experiments. The mid-grain sand commenced with the water-loop module replacing the through air module. For this series of trials, the water-sand tests converged far faster than the air-sand tests reaching steady state in 1-2 hrs. This led to other issues where the temperature of the sand to quickly matched the temperature of the water. Cooling of the sand side was required to ensure the delta in temperature of the water side was large enough to be measureable with a low margin of error. This was remedied by adding ground dry ice at a rate of 1 lb a minute to the sand exiting the discharge funnel onto the conveyor. The dry ice mixed with deposited sand on the conveyor and was further mixed in the elevator ensuring an effective and uniform cooling.

For the experiments heater power chiefly held constant with sand mass flow and air flow being the primarily varied parameters. Some tests where initially run for 3-6 hrs using the construction sand. Subsequent tests and all tests using the mid-grain sand size where run 9-10 hrs to determine a better long term steady measurement. For the water-sand experiments, the tests were run using dry ice for cooling and generally lasted between 1-3 hrs for each test to reach a steady value. The purpose of these tests was to collected data that could yield a comparable heat exchanger performance indicator, represented by UA. A nominal UA was calculated by the following sequence of equations. First the Log Mean Temperature Difference or LMTD across the heat exchanger is calculated as in equation (1):

𝐿𝑀𝑇𝐷 = (∆𝑇top−∆𝑇bot)

ln∆𝑇top∆𝑇bot

(1)

where ∆𝑇top (°C) is the difference between the air or water streams at the top of the heat exchanger with the sand and ∆𝑇bot (°C) is the difference between the air or water streams at the bottom of the heat exchanger with the sand. By taking the heat rate, ��luid (kW), transferred into the system by the working

fluid, in this case either the air or water with equation (2) and dividing by the LMTD a UA can be determined:

��luid = ��luid ∙ 𝐶p ∙ (𝑇in − 𝑇𝑜𝑢𝑡) (2) For the heat rate, ��luid (kg/s) is the mass flow of the

working fluid, either air or water. 𝐶p (kJ/kg-K) is the specific heat of that fluid, 𝑇in (°C) is the measured temperature of the fluid entering the heat exchanger and 𝑇𝑜𝑢𝑡 (°C) is the measured fluid temperature exiting the heat exchanger. The nominal UA, (W/K) of the heat exchanger can then be calculated as:

𝑈𝐴 = 𝐹𝐺 ∙

��luid𝐿𝑀𝑇𝐷

(3) This produced a comparable nominal coefficient of

conductance for the exchanger where FG or exchange effectiveness was assumed to be 1. This assumption is sufficient for quick comparative analysis between test runs, but a more rigorous method to allow wider comparison was followed for the results below. This involved the computational iterating of a procedural calculation method for a multi-pass crossflow exchanger found in Hesselgreaves [19] using the method described by Shah [20]. Likewise a sand side, hsand, and effective heat transfer coefficient, heff, accounting for the fins was calculated to relate a specific heat transfer performance to the exchanger unit. The set of uncertainties of the input measurements are found in Table 3.1.

Table 3.1. Measurement Uncertainties

4. RESULTS Of the sands looked at, course construction sand, mid-grain

silica, and ultra-fine silica, only the mid-grain silica proved suitable. Both the construction sand and ultra-fine silica exhibited too much dust generation over long term use. In the case of the ultra-fine silica, dust generation was so much that running experiments with it was deemed unmanageable. Construction sand was run with air and had results comparable to the mid-grain silica, but the dust generation was such that it was not considered a viable candidate for long term testing or use. Below in Table 4.1 are the results for the mid grain silica, which was of high purity and had little dust generation.

Water Air

Measurement Uxi Units Measurement Uxi Units

T SandIn (1) 0.027 C T SandIn (1) 0.027 C

T SandOut (1) 0.012 C T SandOut (1) 0.012 C

T WaterIn (1) 0.012 C T AirIn (1) 0.043 C

T WaterOut (1) 0.026 C T AirOut (1) 0.017 C

(2) 0.005 gal/min (2) 0.006 ft³/min

T ambient 0.2 F

(1) SRTD Calibration (2) Instrument Error Specs (3) Indiced Reading Error

��Water ��Air


Table 4.1. Air-Sand Runs

Results of the runs show a UA of the heat exchanger at full

flow, ~1 kg/s for mid grain sized sand, at 570-620 W/K with full blower air flow, ~434 SCFM. Significant time was needed for the system to reach a steady state and even 9+ hr runs sometimes yielded a downward tailing in UA near the end. Shop air conditions for the experiment could not be strictly controlled and fluctuated throughout the day, tending to cool the system towards the end of the day. For ~1 kg/s mid grain sized sand flow and ½ air yielded a reduced UA of roughly 370 W/K. For ~0.75 kg/s mid grain sized sand flow and full air flow yielded a lower UA of 515 W/K following the expected trend.

Table 4.2 Water-Sand Runs

The water module test loop had a faster time to steady state generally reaching it in 1-2 hrs, Table 4.2 shows the results. The UA for heat exchanger at full sand flow of ~1 kg/s for mid grain sized sand and 10 GPM water flow was 1390-1465 W/K. The UA at ~0.75 kg/s and 10 GPM was 1290-1385 W/K. The UA for 15 GPM water and ~1 kg/s sand was 1895 W/K. This water module system was much quicker and more stable in delivering a steady set of extended measurements.

5. DISCUSSION Of the two sets of test results the water-sand runs would be

the most reliable, as they all tended to reach a definite steady state; representative runs are shown in Fig. 5.1. The proportional error propagation also indicates a stronger certainty in hsand results for the water as well. Furthermore the data from water operation is more reliable since the sand-side resistance dominates when water is the heat source. The results in Table 4.2 show full flow sand side heat transfer coefficient (heff) of 200-275 W/m2-K when related to the base area, meaning the outer tube area. In contrast, the coefficient (hsand) is only 12 to 17 W/m2-K when referred against the total effective fin area. This is corrected for the very high fin efficiency carefully calculated to apply in this case. The observed sand side coefficient is less than the targeted 100 W/m2-K, which is approximately the value observed in smaller lab-scale heat transfer tests conducted at Georgia Tech. The spiral finned tube is relatively inexpensive, so it is not unreasonable to consider it as more of a surface enhancement rather than as true extended surface as in a longitudinal finned tube. From this point of view, the observed performance is adequate; nevertheless, the higher 100 W/m2-K sand side effectiveness was a performance target based on lab-scale results, which has not yet been demonstrated. We believe that we understand the reasons for this shortcoming and will beworking to alleviate this problem. Likely causes of this are the extremely slow sand flow speeds of 0.6mm/s-0.9mm/s which were a restriction of what the then available OLDS Elevator model could deliver being undersized for the full-scale test component. Another contributing factor could be an observed unevenness in distribution of flow through the heat exchanger at the top, particularly the corners, potentially leading to stagnant zones on the heat exchanger walls. This would reduce the effective heating area of the exchanger.

6. CONCLUSION The lower than expected sand side heat transfer coefficient

is caused in part by the lower sand speed in this study. A speed of 10 mm/s giving around 100 W/m2-K could easily be achieved in practice rather than the experimental speed of less than 1 mm/s. Extrapolation of the lab data to lower speed shows that the coefficient should increase at least 50% with such an increase in speed, which is helpful but not a sufficient explanation. The more likely explanation is poor sand flow distribution among the tubes and between the fins. The tested serpentine particle flow heat exchanger has tubing with 10 fins per inch (FPI), which was the widest spacing found to be available for bent-to-shape serpentine tubing. In contrast, the experimental tubing has 8 FPI, which provided better flow between tubing. In fact, the laboratory tests show about the same sand side heat transfer coefficient for bare tubes as 8 FPI tubes demonstrating good fin to sand contact with such tubes. Poor flow between bends is likely exacerbated in the bends, so and fin contact issues could easily account for a 1.5 x 2 = 3 times differenced in observed heat transfer coefficient.

Sand-Air Units Mid-Grain Silica

ConditionsFull flow,

Full Blower0.75 flow,

Full BlowerFull flow,

Half Blower

UA 570-620 370 515Uncertainty ±12 ±6 ±6

h sand 8.4-9.7 5.8 6.8Uncertainty ±0.4 ±0.2 ±0.1

h eff 140-165 98.1 114.7Uncertainty ±6.5 ±3.3 ±2Sand Speed (mm/s) 0.6-0.9 0.5-0.7 0.6-0.9

(W/m²-K)

(W/m²-K)

(W/K)

Sand-Water Units Mid-Grain Silica

ConditionsFull flow,

Full Blower0.75 flow,

Full BlowerFull flow,

Half Blower

UA 1390-1465 1290-1385 1895Uncertainty ±45 ±33 ±49

h sand 12.4-13.3 11.3-12.3 17.2Uncertainty ±0.5 ±0.4 ±0.8

h eff 200-215 185-200 275Uncertainty ±8 ±6 ±11Sand Speed (mm/s) 0.6-0.9 0.5-0.7 0.6-0.9

(W/m²-K)

(W/K)

(W/m²-K)


Figure 5.1. Test Run Stability


It is also possible that the sand flow is poorly distributed among the tubes. In fact, when the heat exchanger was originally tested it was seen that a rat-hole funnel flow was present, which likely would have reduced the effective heat transfer area to 25% of the total area. This obvious funnel flow problem was eliminated by an empirically developed baffle installed in the exit funnel. We were suspicious of non-uniform flow when the low UA data were measured, and attempted with some success to visualize the flow by an innovative measure of installing bore scope probes into the funnel. Some definite irregularity in the flow favoring about one half of the exchanger over the other half was evident. Current development is a suitable redesigned of the heat exchanger to install back into an improved test apparatus updated with a new higher flow rate OLDS elevator to achieve 100 W/m2-K for the hsand.

NOMENCLATURE

𝐶p Specific heat of working fluid FG Exchange Effectiveness Coefficient FPI Fins per Square Inch heff Effective heat transfer coefficient hsand Sand aide heat transfer coefficient LMTD Logarithmic mean temperature difference ��luid Mass flow rate of working fluid RTD Resistance Temperature Detector SRTD Standardized Resistance Temperature Detector SMD Sauter Mean Diameter 𝑇in Temperature of fluid entering heat exchanger 𝑇out Temperature of fluid exiting heat exchange ∆𝑇bot Temperature difference of enter and exit working fluid streams at bottom of exchanger ∆𝑇top Temperature difference of enter and exit working fluid streams at top of exchanger TAmbient Ambient air temperature TAirIn Air inlet temperature TAirOut Air outlet temperature TSandIn Sand inlet temperature TSandOut Sand outlet temperature TWaterIn Water inlet temperature TWaterOut Water outlet temperature UA Conductance of Heat Exchanger ��luid Heat rate of working fluid ��Air Volumetric flow rate of air ��Water Volumetric flow rate of water

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