Performance study of the Maisotsenko Cycle heat exchangers in different air-conditioning applications PDF

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International Journal of Heat and Mass Transfer 81 (2015) 207–221 Contents lists available at ScienceDirect International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt Performance study of the Maisotsenko Cycle heat exchangers in different air-conditioning applica...

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International Journal of Heat and Mass Transfer 81 (2015) 207–221

Contents lists available at ScienceDirect

International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt

Performance study of the Maisotsenko Cycle heat exchangers in different air-conditioning applications Demis Pandelidis a,⇑, Sergey Anisimov a, William M. Worek b a b

´ ski st., 50-370 Wroclaw, Poland Department of Environmental Engineering, Wroclaw University of Technology, 27 Wyspian Stony Brook University, College of Engineering and Applied Sciences, Department of Mechanical Engineering, 127 Engineering Bldg., Stony Brook, NY 11794, USA

a r t i c l e

i n f o

Article history: Received 12 August 2014 Received in revised form 13 October 2014 Accepted 13 October 2014

Keywords: Maisotsenko Cycle M-Cycle Indirect evaporative cooling Mathematical model Cross-flow recuperator

a b s t r a c t This paper investigates a mathematical simulation of the heat and mass transfer in the two different Maisotsenko Cycle (M-Cycle) heat and mass exchangers used for the indirect evaporative cooling in different air-conditioning systems. A two-dimensional heat and mass transfer model is developed to perform the thermal calculations of the indirect evaporative cooling process, thus quantifying the overall heat exchangers’ performance. The mathematical model was validated against the experimental data. Numerical simulations reveal many unique features of the considered units, enabling an accurate prediction of their performance. Results of the model allow for comparison of the two types of heat exchangers in different applications for air conditioning systems in order to obtain optimal efficiency. Ó 2014 Elsevier Ltd. All rights reserved.

1. Introduction In recent years, the increase in summer temperatures, improved insulation of the buildings, and a growth of indoor facilities have led to an increased requirement for air conditioning in buildings. Conventional mechanical vapor–compression air-conditioning systems consume a large amount of the electrical energy that is largely dependent upon a fossil fuel. This mode of air conditioning is, therefore, neither sustainable nor environmentally-friendly. Due to the increasing need for air conditioning and the growing interest in energy savings, seeking ways to reduce fossil fuel consumption and to increase usage of the renewable energy during airconditioning process in building sector is a matter of great importance. Evaporative air cooling is an alternative to the conventional vapor–compression systems to meet above mentioned economic, environmental, and regulatory challenges. Direct evaporative cooling is the process of evaporating liquid water into the surrounding air and causing its temperature to decease. A typical direct evaporative cooler uses a fan to draw in outside air through a pad-wetting media and circulates the cool air through the building. Theoretically, the ultimate temperature for the direct evaporative cooling process is the ambient air wet-bulb temperature, however this temperature is not easy reached and the resulting air stream is

⇑ Corresponding author. E-mail address: [email protected] (D. Pandelidis). http://dx.doi.org/10.1016/j.ijheatmasstransfer.2014.10.033 0017-9310/Ó 2014 Elsevier Ltd. All rights reserved.

humid. Therefore, new methods and technologies are needed for cooling of buildings. One of the best solutions to this limitation is the sub-wet bulb temperature evaporative cooling. There are several studies on achieving sub-wet bulb temperatures by the evaporative cooling and some innovative ideas do exist. Stoichkov [1] presented a mathematical model describing a cross-flow heat exchanger with a flowing down water film. In this study the wet-bulb temperature of an ambient air was not reached at the exit. Ren and Yang [2] developed an analytical model for the coupled heat and mass transfer processes in an indirect evaporative cooling with parallel/counter-flow configurations. Maclaine-Cross and Banks [3] referred to that for regenerative evaporative cooling the process can approach the dew point temperature of the ambient air if appropriate mass flows and cooler geometry are chosen. Hasan [4,5] numerically simulated various configurations of indirect evaporative coolers. The results showed that the performance of the system could be improved by manipulating the air flow by branching the working air from the product air, which is indirectly pre-cooled. Zhao et al. [6] presented a numerical study of the counter-flow heat and mass exchanger for the dew point evaporative cooling purpose. They proposed a range of design conditions and flow rates to improve the cooler performance. Wang [7] studied the effect of the wettability (the surface wettability factor is the parameter used to estimate the effect of an incomplete wetting) of aluminum plates to the cooling performance of the indirect evaporative systems. A dynamic contact analyzer was applied to

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Nomenclature cp d F h H G L, l M NTU q qo Q b Q

RH t t v W x X Y

specific heat capacity of moist air [J/(kg K)] hydraulic diameter [m] surface area [m2] specific enthalpy of the moist air [kJ/kg] height [m] moist air mass flow rate [kg/s] streamwise length of cooler [m] water vapor mass transfer rate [kg/s] number of transfer units, NTU ¼ aF=ðGcp Þ [–] heat flux [W/m2] latent heat water [kJ/kg] rate of heat transfer [W] specific cooling capacity per cubic meter of the heat exchanger’s structure [kW/m3] relative humidity [%] temperature [°C] average temperature [°C] air stream velocity [m/s] heat capacity rate of the fluid [W/K] humidity ratio [kg/kg] coordinate along primary air flow direction [m] coordinate along working air flow direction in the wet channels [m]

Special characters a convective heat transfer coefficient [W/(m2 K)] b mass transfer coefficient [kg/(m2 s)] d thickness [m] k thermal conductivity [W/(m K)] e thermal effectiveness [%] r surface wettability factor, r 2 (0.0–1.0) [–] Non dimensional coordinates Le Lewis factor Le ¼ a=ðbcp Þ [–] NTU number of transfer units NTU ¼ aF=ðGcp Þ [–]

quantitatively measure the advancing and receding contact angles and the water-retention capacity of different aluminum surfaces. Riangvilaikul and Kumar [8] carried out experimental studies on a dew point evaporative cooling exchanger. Their results indicated that the wet-bulb effectiveness achieved by the exchanger was 92– 114%. Zhou [9] carried out a study to optimize the design of the water distributor to improve the water distribution uniformity in the indirect evaporative air coolers. This research outlined several available water distribution modes applicable to these types of devices. To enhance the cooling performance of the typical evaporative exchangers, a novel thermodynamic cycle, known as the M-Cycle [10–18], was proposed by Professor Valeriy Maisotsenko as the new approach of making and operating the heat and mass exchanger (HMX). This was claimed to enable harnessing extra amount of energy from the ambient using a dedicated flat plate, cross-flow and perforated heat exchanger. According to the producer, the maximal water consumption for the unit containing six HMXes is 14 kg/h [13,17]. The water needed to provide 1 kW of cooling is only 0.001 kg (for one HMX, at inlet conditions ti = 30 °C and RHi = 40% and primary and working air flow rate equal 330 m3/h [17]). In the typical indirect evaporative air coolers (Fig. 1(a)) the working air flow is delivered directly to the wet channel, while the primary air flow is delivered to the dry channels. The principal idea of the Maisotsenko Cycle is to indirectly pre-cool the working airflow before it is delivered to the wet channel (Fig. 1(b)). Passing

Nu Pr Re X Y

Nusselt number [–] Prandtl number [–] Reynolds number [–] X ¼ X=l – relative X coordinate [–] Y ¼ Y=l – relative Y coordinate [–]

Subscripts 1 main (primary) air flow 2 working (working) air flow in the wet channels (product part of exchanger) 3 working (working) air flow in the dry channels (pre-cooling part of exchanger) 4 working (working) air flow in the wet channels (pre-cooling part of exchanger) cond heat transfer by thermal conduction Icond referenced to the first-order boundary conditions IIcond referenced to the second-order boundary conditions g water vapor h referenced to the height of the channel heat heat transfer i inlet l latent heat flow mass mass transfer o output p plate surface plt channel plate product referenced to the product section of heat exchanger s sensible heat flow w water film work referenced to the working section of heat exchanger WB wet-bulb temperature X air streamwise in the dry channel Y air streamwise in the wet channel  referenced to the elementary plate surface 0 condition at the air/plate interface temperature

through the dry channel the primary air flow is cooled without increasing the moisture content (process 1i–1o). At the outlet of the dry channel a part of the primary flow (working air flow) is delivered to the wet channel, where it realizes the evaporative cooling process on the base of direct air–water contact (process 2i–2o in Fig. 1(b)). The remaining portion of the primary flow (product air flow) is delivered to the conditioned space. The currently produced HMX has a unique design to maximize the efficiency of the direct and indirect stages of cooling process. Fig. 1(c) illustrates air-flow arrangement in the HMX produced by the Coolerado Corporation [10]. The working mechanism of the HMX under the M-Cycle is described as follows. Part of the surface on the dry side is used for the primary air flow (Stream 1 in Fig. 1), while the rest is used for the working air flow (Stream 2 in Fig. 1). The primary and the working air streams are guided to flow over the dry side along parallel flow channels. The working air stream is delivered to the dry channels first, in order to be sensibly precooled before it is split into multiple streams that are directed into the wetted section. There are regularly distributed holes in the channel where the working air is retained and each of these allows a certain percentage of the air stream to pass through the wet channels. The working air stream is then gradually delivered to the wet channel (Stream 3 in Fig. 1) as it flows along the dry side, forming an even distribution of the air flow over the wetted surface. The pre-cooled working air, delivered to the wet channel, flows over the wet

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Fig. 1. The principal idea of the M-Cycle and the novel heat and mass exchanger which utilizes the cycle: (a) principal operation of the typical indirect air cooler on the psychrometric chart; (b) principal operation of the air cooler on the base of the M-Cycle on the psychrometric chart; (c) currently produced M-Cycle HMX: 1 – primary air flow; 2 – working air flow (dry channel); 3 – working air flow (wet channel).

surface arranged at right angles to the dry-side channels, absorbing heat from the working air flow in the dry channels and the primary air stream. Owing to its pre-cooling effect, the working air in the wet side has a lower temperature and therefore, it is able to assimilate more heat from the primary flow. As a result, the wet-bulb effectiveness of the novel M-Cycle unit is significantly higher than the effectiveness of the traditional cross-flow exchangers (see Refs. [11,12]). From the literature, it can be found that the M-Cycle has been investigated by several research groups. Gillan [10] presented the study of an applied M-Cycle unit for the commercial purposes. Worek et al. [14] showed the possibility of using the M-Cycle HMX in desiccant-indirect evaporative air-cooling system for various

ambient conditions. Zhan et al. [12] conducted numerical analyses of the M-Cycle including a cross-flow heat exchanger to obtain thermal performance. A finite-element method was used with experimental data for comparison purposes. Miyazaki [15] analyzed an M-Cycle for an integrated air-cooling system driven by the solar energy for air conditioning. As shown above, the interest in the M-Cycle is growing and the new cycle has a potential to replace a significant part of the mechanical compression system load. However, no studies have been conducted that focus on the comparative analysis of the different applications of the M-Cycle heat and mass exchanger in different air-conditioning systems. In this regard, this study focuses on the comparison of two types of

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the M-Cycle cross-flow HMXs for different applications in air-conditioning systems. The HMXs considered are shown schematically in Fig. 2. The first device (HMX1 – Fig. 2(a)) is currently produced by Coolerado Corporation, where the working and the primary air flows in the dry channels are flowing parallel to each other. The second unit (HMX2 – Fig. 2(b)) is the modification of the first cross-flow HMX, where entrance to the dry channel for the working and the product air flow are placed on the opposite sides of the HMX, which allows for implementation in different configurations in air-conditioning systems. Both units can be used as an individual cooling coils, which are the only source of cooling power in the system. However, in many air conditioning systems, especially in the public buildings and offices, the individual rooms typically use fan coil units to provide individual comfort for the occupants (Fig. 3(a)–(d)). The exhaust air for the systems with the original HMX is removed to the outside or it is partly recirculated to the primary air flow (Fig. 3(a)). The modified HMX allows for the system to be implemented in the supply–exhaust flow stream in the air-handling unit (Fig. 3(b)). Therefore it can operate on the exhaust air, which is colder than the ambient air. The system presented in Fig. 3(b) can be replaced with a system having a heat pump, presented in Fig. 3(c). The operation base in this case is the same as in supply–exhaust system with the fan coil units. In more humid climates, there is a possibility to use the original HMX with the desiccant wheel (Fig. 3(d)), which dries the air flow and increases its temperature [14]. The arrangements presented

require different operational conditions for the presented HMXs, with different temperatures and relative humidities of the primary and the working air-flow entering each exchanger. It is essential to establish the conditions for which it is more reasonable to use one HMX instead of another. In this paper such an evaluation will be conducted using a numerical model.

2. Methods In this paper an analytical model based on the modified e–NTU method is developed to analyze the performance of two M-Cycle HMXs. In the e–NTU-model, the air stream in matrix passages is considered as a gaseous fluid flow with constant temperature, velocity and mass transfer potential (humidity ratio of air flow) in the direction normal to the plate surfaces, which equal to bulk average values [11,16–18]. The e–NTU method was used to describe indirect evaporative air coolers with satisfactory agreement with the experimental results presented by Hasan [5] and Miyazaki [15]. The authors developed the original e–NTU-model describing coupled heat and mass transfer in considered M-Cycle heat exchangers. Fig. 4 presents the assumed air flow distribution inside considered devices, with basic information about the inlet conditions. From the standpoint of the heat and mass transfer processes occurring in the channels of considered M-Cycle HMXs can be divided into two main sections [11]. The first one is a typical

Fig. 2. Analyzed HMXs: (a) original cross-flow HMX (HMX1); (b) modified cross-flow HMX (HMX2).

Fig. 3. Analyzed heat exchangers in different applications for air conditioning systems: (a) HMX1 in air conditioning system with fan coils; (b) HMX2 in air conditioning system with fan coils; (c) HMX2 in air conditioning system with heat pump; (d) HMX1 in air conditioning system with desiccant wheel.

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Fig. 4. Air flow distribution in considered HMXs: (a) HMX 1; (b) HMX 2.

cross-flow indirect evaporative cooler (this portion is responsible for cooling the primary air flow), and the second one, the task of which is pre-cooling the working air flow and formation effective temperature distribution of the working air flow at the entrance to the wet channels of the first section of the heat exchanger (Fig. 4). In this regard, depending on the analyzed section, the working air stream in both devices will be considered as three different flows 2, 3, 4 (Fig. 4). In the first section, the pre-cooled working airflow 2 is exchanging heat with the primary airflow 1, like in an ordinary cross-flow indirect evaporative cooler but with nonuniform initial distribution of the working air flow’s temperature and humidity ratio at the entrance to the channel 2. The second and third portions are responsible for pre-cooling the working air. The above-mentioned portion of the working air stream will be marked by quantitatively equal 2, 3 and 4, where 2 is the part which exchanges heat with product air 1, 3 is pre-cooling air flow which does not contact water film (working air in the dry channel) and 4 is pre-cooling air flow which contacts the water film. The balance equations for air flow 3 consider only sensible heat transfer, while set of equations for air flow 4 describes combined heat and mass transfer. Stream 3 and 4 are mixed in the wet channel (Fig. 4). The proposed mathematical model for considered HMXs was based on the following assumptions:  Heat losses to the surroundings are negligible.  Steady state operation.  Driving force of mass transfer is a gradient of moisture content (vapor’s partial pressure).  Airflow is an ideal, incompressible gas mixture of dry air and water vapor.  Longitudinal molecular diffusion of water vapor in air and longitudinal heat conduction along the wall as well as inside the fluids in the direction of air flow are negligible.  Kinetic properties of air flow and water film are constant and equal to bulk average values.  Consumed water rate corresponds to sufficient evaporation and keeping up the material of plates in hygroscopic saturated condition. This causes that air flow heat capacity to be much larger than that of the water film (i.e., W2  Ww.

 The passage walls are impervious to mass transfer.  The temperature of the water film, the sensible heat transfer coefficient a and the Lewis factor depend on the operating conditions [11]. 2.1. Mathematical model equations In this section, the equations describing the product part of considered HMXs will be presented, equations describing the initial portion of the exchanger are analogous, therefore they will be omitted. The only difference in mathematical description of processes occurring in considered parts is the fact, that equations describing the initial part additionally consider an algorithm that describes air streams mixing (which was presented by authors in [11]) and must include the variation of the air Stream 3 and 4, due to the continuous mixing of those streams in the wet channel. The equations describing the product part are identical for HMX1 and 2, where the difference lies in the different initial conditions. According to the above assumptions the following heat and mass balance equations can be written for the air streams passing through control volume in the product part of the HMX (see Fig. 4 and 5)  For the air stream in the dry channel. o The energy conservation balance considering only sensible heat transfer on the plate surface (see Fig. ...