Airside Thermal and Hydraulic Performance of a Bare Tube Heat Exchanger with Diameter of 0.8 mm Under Dehumidifying Conditions.
Lots of research has been done to enhance the performance of air-to-refrigerant heat exchangers (HX) due to its significance in air-conditioning and heat pump systems. Conventional heat exchangers, such as fin-and-tube and microchannels, use fins to augment the heat transfer area of the air-side. Various fin designs are widely studied and are available in the literature (Wang et al., 2000, Singh et al., 2009, 2011), including slit fins (Wang et al., 1999), fins with longitudinal vortex-generator (Kwak et al., 2002), and crimped spiral fins (Tang et al., 2009). However, the tube diameter range for these fin-and-tube heat exchangers are all larger than 5 mm.
Recently researchers (Bacellar et al., 2016a; Bacellar et al., 2016b; Shabtay et al., 2016) found that finless designs using [less than or equal to] 1 mm hydraulic diameter bare tubes can outperform air-side heat transfer than heat exchangers with fins under dry conditions. Bare tube heat exchanger has two times larger air side convective heat transfer coefficient than state-of-art microchannel heat exchangers under dry conditions (Shabtay et al., 2016). However, there is no research on heat, mass transfer and friction characteristics on such novel heat exchanger under dehumidifying condition yet.
This paper presents the experimental investigation of the thermal and hydraulic characteristics of a novel bare tube heat exchanger prototype, manufactured by using stainless steel tubing with outer diameter of 0.8 mm, under wet condition. It was tested both vertically and horizontally.
PROTOTYPE AND TEST MATRIX
This novel bare tube heat exchanger (BTHX) was manufactured using stainless steel tubes with an outer diameter of 0.8 mm and inner diameter of 0.6 mm, as shown in Figure 1 (a). The vertical and horizontal orientation are shown in Figure 1 (b) and (c). The detailed manufacturing method for fabricating the prototype can be found in Shabtay et al., 2016. Frontal area of the BTHX is 0.0228 [m.sup.2], and air-side heat transfer area is 0.1826 [m.sup.2].
Though the focus of this paper is to present the thermal and hydraulic characteristics of this heat exchanger in wet mode, dry condition test results are also shown to better explain the effect of inlet air humidity. This heat exchanger prototype was tested under both vertical and horizonal orientations, as summarized in Table 1. The inlet air conditions are fixed dry bulb temperature of 26.7[degrees]C with varous relative humidities of 35%, 50% and 70%. The inlet air frontal velocity varies at 3, 6 and 9 m/s. The inlet water temperature is 12[degrees]C and the water mass flow rates are 20, 35 and 50 g/s, respectively. For the horizontal orientation test, only conditions of the smallest and largest air velocity were tested.
For the dry conditions, Wilson plot method is used to calculated air-side heat transfer coefficient (AHTC) (Wilson, 1915). Wilson plot method is a widely adopted method to determine the convective heat transfe coefficient (HTC) using experimental data. Detailed procedures can be found in Wilson, 1915.
For the wet conditions, Threlkeld (1970) method was employed. The main steps are summarized here. The total heat transfer rate for air side is shown in Equation (1). The water side heat transfer rate is shown in Equation (2). The overall heat transfer coefficient [U.sub.ow], based on the on the enthalpy potential is given in Equation (3).
[Q.sub.a,w] = [m.sub.a] ([i.sub.a,o] - [i.sub.a,i]) (1)
[Q.sub.w] = [C.sub.p,w][m.sub.w] ([T.sub.w,i] -[T.sub.w,o]) (2)
[Q.sub.ave,w] =[U.sub.o,w] [A.sub.o] [DELTA][i.sub.m] F (Error!
Bookmark not defined.3)
where [DELTA][i.sub.m] is the mean enthalpy difference for counter flow coil, defined as in Equation (4), (5) and (6)
[DELTA][i.sub.m] = [i.sub.a,m] - [i.sub.r,m] (Error!
Bookmark not defined.4) [mathematical expression not reproducible] (5)
[mathematical expression not reproducible] (Error!
Bookmark not defined.6)
The overall heat transfer coefficient is related to the individual heat transfer resistance (Myers, 1967) as follows:
[mathematical expression not reproducible] (7)
Tube-side heat transfer coefficient, [h.sub.i] is evaluated from the Gnielinski correlation (Gnielinski, 1976). The calculation procedure to determine sensible heat transfer coefficient [h.sub.c,o] and mass transfer coefficient [h.sub.d,o] can be found in Pirompugd et al. (2005). Air-side heat and mass transfer performance and friction under wet condition are evaluated using Chilton-Colburn j, Chilton-Colburn [j.sub.m] and f factor. Chilton-Colburn [j.sub.m] is as follows:
[mathematical expression not reproducible] (8)
[mathematical expression not reproducible] (9)
[mathematical expression not reproducible] (10)
RESULTS AND DISCUSSION
The results from the wet test conditions are summarized in Figure 2 through Figure 11, where vertical orientation is on the left and horizontal orientation is on the right. Energy balance (EB) for all data points are withn [+ or -] 5%, as shown in Figure 2 and Figure 3. Here we discuss the effects of inlet air humidity, air flow rate and water flow rate and orientation on heat exchanger capacity, sensible heat, latent heat and airside pressure drop (ADP), respectively.
Effect of inlet air relative humidity (RH)
When inlet air relative humidity (RH) is 35%, the heat exchanger is at dry condition, meaning there is only sensible cooling. As inlet air relative humidity increases from 35% to 70%, overall heat transfer capacity increases (Figure 4) for the vertical orientation tests. However, when air RH changes from 35% to 50%, heat exchanger capacity decreases slightly at horizontal orientation (Figure 5). Increased RH also leads to lower sensible heat (SH), as shown in Figure 6 and Figure 7. This is because higher inlet air humidity leads to additional condensing water accumulation on the heat exchanger surface, which reduces dry surface area and restrains sensible heat transfer. Accordingly, latent heat increases (Figure 8 and Figure 9) as inlet RH increases. In terms of airside pressure drop, larger inlet air humidity results in larger airside pressure drop due to the bridging effect formed by retained condensate water between the tubes (Figure 10 and Figure 11).
Effect of inlet air flow rate (AFR)
As air flow rate (AFR) increases, total capacity (Figure 4 and Figure 5), sensible capcity (Figure 6 and Figure 7) and airside pressure drop (Figure 10 and Figure 11) all increase while latent heat may increase or decrease (Figure 8 and Figure 9). The change of latent heat is also affected by other factors such as inlet air humidity, water flow rate, condensate removel, heat exchanger orientation and heat exchanger geometry. For this HX, the latent heat decreases may increase or decrease depending on test condition (Figure 8 and Figure 9). Generally, as air flow rate increases, total capacity increases, causing water outlet temperature to increase, resulting in higher average wall temperature. The latent heat transfer is expected to decrease due to increased wall surface temperature. But larger air flow rate also means more moisture in the airstream, which produces more condensation. Besides, there are also other factors need to be considered such as the ease of condensate water removal. Thus whether latent heat increases or decreases depends on which factor is dominant. Here we first explain the effect of AFR under vertical orientation. The most important factor under vertical orientation is the inlet air RH. When inlet air RH is low, the surface is partially wet and sensible heat transfer is the dominant factor, leading to latent heat decrease. This phenomenon is observed in the experiment, as shown in Figure 12 (a) and (b). Compare the results at AFR=0.06 and 0.09 [m.sup.3]/s, it can be noticed that as air flow rate increases, wet surface area becomes smaller and dry surface area becomes larger, especially on the top of the heat exchanger. When inlet air RH is high, the surface is fully wet, the extra moisture in the air becomes the dominant factor, causing latent heat to increase. Figure 12 (c) and (d) show that both surfaces are fully wet. Next, we explain the effect of AFR under horizontal orientation. Here, the most crucial factor is removal of condensate water. At horizontal orientation, it is hard to remove the condensate water at low flow rate due to the orientation, as shown in Figure 13 (a). Instead of flowing along the tubes as in the cases of vertical orientation, condensate water just accumulates in between the tubes and form a water bridge if the air flow is not strong enough. At higher velocity as shown in Figure 13 (b), the condensate water is blown out by the incoming air flow. Water splashes in the downstream of air flow, leaving water marks on the wind tunnel duct wall. For pressure drop, higher air flow rate results in higher ADP, which is expected (Figure 10 and Figure 11). One interesting phenomenon is under vertical condition, airside pressure drop for RH=35% and 70% is not much affected by waterside flow rate, but airside pressure drop for RH=50% increases as water flow rate increases. This is because for each heat exchanger geometry, there is a certain maximum amount of water retention corresponding to a specific condition. When RH is 50%, the amount of retained water has not reached its maximum yet, thus the increase of waterside flow rate will cause latent heat increase, enhancing the bridging effect. When RH is 70%, the amount of retained water has reached maximum, and consequently the extra condensing water flows down the tubes.
Effect of inlet water flow rate (WFR)
As water flow rate increases, total capacity (Figure 4 and Figure 5), sensible heat (Figure 6 and Figure 7) and latent heat (Figure 8 and Figure 9) all increase. Water flow rate increase has a negelible effect on airside pressure dorp when inlet air RH=35% and 70% but it leads to airside pressure drop increase when RH=50%. This has been discussed previously.
Effect of heat exchanger orientation
When there is no water condensate, there is no difference regarding total capacity, sensible heat, latent heat and airside pressure drop between heat exchanger operating under either orientation. Under the wet conditions, compared with heat exchanger under vertical orientation, the horizontal orientation tests show smaller total capacity (Figure 4 and Figure 5), sensible heat (Figure 6 and Figure 7) and latent heat (Figure 8 and Figure 9). However, the airside pressure drop (Figure 10 and Figure 11) under horizontal condition is larger than that under vertical condition due to bridging effect. Thus we recommend to operate this heat exchanger under vertical condition if there is condensation.
j, jm and f factor
Here we only discuss the j, [j.sub.m] and f factor results under the preferred vertical condition. As it was already demonstrated that existing j and f factor correlations for bare tube bundles under dry condition in literature are not applicable for such small diameter tubes (Bacellar et al., 2016a) and there is no available wet condition correlations, new correlations are developed.
Figure 14 presents the variation of f factor on the effect of inlet air RH and Reynolds number. Friction increases as inlet air RH increases from 35% to 70% due to the bridging effect of condensate water. Correlations of f factor are developed as power law based on Reynolds number, as shown in Equation (11), (12) and (13). Maximum deviation for f factor is within [+ or -]2% under dry condition and [+ or -]15% under wet condition (Figure 15).
[mathematical expression not reproducible] (11)
[mathematical expression not reproducible] (12)
[mathematical expression not reproducible] (13)
Figure 16 and Figure 18 presents the variation of j and [j.sub.m] factor on the effect of inlet air RH and Reynolds number. The variation of j and [j.sub.m] factor on the relative humidity effect of inlow air is not sensitive. This is consistent with observations by Wang et al., (2000) and Phan et al., (2011). Correlations of j and [j.sub.m] factor are developed as power law based on Reynolds number, as shown in Equation (14) and (15). Maximum deviations for j and [j.sub.m] factor are within [+ or -] 10% (Figure 17 and Figure 19).
[mathematical expression not reproducible] (14)
[mathematical expression not reproducible] (15)
In current study, the heat and mass transfer characteristics of a novel compact heat exchanger consisting of round bare tubes with diameter of 0.8 mm are studied under dehumidifying conditions. The effect of inlet air humidity, air flow rate, water flow rate and orientation (vertical and horizontal) of heat exchanger were discussed. Experimental results show that total heat transfer capacity increases as inlet air humidity increases, air flow rate increases and water flow rate increases. The change of latent heat is also affected by other factors such as inlet air humidity, water flow rate, heat exchanger orientation and heat exchanger geometry. Airside pressure drop increases as air flow rate increases and inlet air humidity increases. Water flow rate increase has negligible influence when the surface is total wet or total dry and leads to pressure drop increase when surface is partical wet. With increasing Reynolds number, j, [j.sub.m] and f factor decrease. f factor increases as air humidity increases due to the formation of condensate and bridging effect. The effect of inlet air humidity on j and [j.sub.m] factors are not obvious. Chilton-Colburn j, [j.sub.m] and f factor power law correlations were developed for Reynolds number range of 200~600. Maximum deviations are [+ or -]15%, [+ or -]10% and [+ or -]10%, respectively.
The authors gratefully acknowledge the support of this effort from the Center for Environmental Energy Engineering (CEEE) at the University of Maryland.
A = area ([m.sup.2])
[b'.sub.p] = slope of a straight line between the outside and inside tube wall temperature (J*[kg.sup.-1]*[K.sup.-1])
[b'.sub.r] = slope of the air saturation curved at the mean coolant temperature and the inside wall temperature (J*[kg.sup.-1]*[K.sup.-1])
[b'.sub.w,p] = slope of the air saturation curve at the water film temperature of the tube wall surface (J*[kg.sup.-1]*[K.sup.-1])
[b'.sub.w,m] = slope of the air saturation curve at the water film temperature of the fin surface (J*[kg.sup.-1]*[K.sup.-1])
j = Chilton-Colburn j factor for heat transfer
[j.sub.m] = Chilton-Colburn j factor for mass transfer
f = airside friction factor
h = convective heat transfer coefficient (W*[m.sup.-2]*[K.sup.-1])
i = enthalpy (kJ*[kg.sup.-1])
k = thermal conductivity (W*[m.sup.-1]*[K.sup.-1])
Le = Lewis number
[L.sub.p] = tube length (m)
[??] = mass flow rate (kg*[s.sup.-1])
Q = heating capacity (W)
Sc = Schmidt number
T = temperature (K or [degrees]C)
[U.sub.o,w] = wet surface overall heat transfer coefficient, based on enthalpy difference (kg*[m.sup.-2]*[s.sup.-1])
[rho] = density (kg*[m.sup.-3])
[sigma] = contraction ratio
a = air side
ave = average
i = inlet
m = mean
o = outlet, overall
p = tube wall
w = water side, wet
fr = frontal
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Bacellar, D., Aute, V., Huang, Z., Radermacher, R. (2016b). Novel Air-side Heat Transfer Surface Designs Using an Integrated Multi-Scale Analysis with Topology and Shape Optimization. International Refrigeration and Air Conditioning Conference. Paper 1610.
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Table 1. Test matrix Inlet Air Temperature [[degrees]C] Inlet Air RH [%] Inlet Water [[degrees]C] Temperature Air Flow Rate [m3/s] 0.03 (Vertial & Horizontal) Water Flow Rate [g/s] 20 35 50 Inlet Air 26.7 Temperature Inlet Air RH 35 (dry), 50 (wet), 70 (wet) Inlet Water 12.0 Temperature Air Flow Rate 0.06 (Vertial) Water Flow Rate 20 35 50 Inlet Air Temperature Inlet Air RH Inlet Water Temperature Air Flow Rate 0.09 (Vertial & Horizontal) Water Flow Rate 20 35 50
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|Author:||Huang, Zhiwei; Ling, Jiazhen; Hwang, Yunho; Radermacher, Reinhard|
|Publication:||ASHRAE Conference Papers|
|Date:||Jan 1, 2018|
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