Experimental analysis of the thermal entrainment factor of air curtains in vertical open display cabinets for different ambient air conditions

Applied Thermal Engineering 31 (2011) 961e969

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Experimental analysis of the thermal entrainment factor of air curtains in vertical open display cabinets for different ambient air conditions Pedro Dinis Gaspar a, *, L.C. Carrilho Gonçalves a, R.A. Pitarma b a

University of Beira Interior, Electromechanical Engineering Department, Rua Fonte do Lameiro e Edifício 1 das Engenharias, 6201-001 Covilhã, Portugal Polytechnic Institute of Guarda, High School of Technology and Management, Mechanical Engineering Department, Avenida Dr. Francisco Sá Carneiro, n
50, 6300-559 Guarda, Portugal b

a r t i c l e i n f o

a b s t r a c t

Article history: Received 25 November 2009 Accepted 17 November 2010 Available online 24 November 2010

The vertical open refrigerated display cabinets suffer alterations of their thermal performance and energy efficiency due to variations of ambient air conditions. The air curtain provides an aerothermodynamics insulation effect that can be evaluated by the thermal entrainment factor calculation as an engineering approximation or by the calculus of all sensible and latent thermal loads. This study presents the variation of heat transfer rate and thermal entrainment factor obtained through experimental tests carried out for different ambient air conditions, varying air temperature, relative humidity, velocity and its direction relatively to the display cabinet frontal opening. The thermal entrainment factor are analysed and compared with the total sensible and latent heats results for the experimental tests. From an engineering point of view, it is concluded that thermal entrainment factor cannot be used indiscriminately, although its use is suitable to design better cabinet under the same climate class condition. Ó 2010 Elsevier Ltd. All rights reserved.

Keywords: Thermal entrainment factor Refrigerated display cabinet Experimental data

1. Introduction A large part of the refrigeration equipments installed in supermarkets are vertical open refrigerated display cabinets with a frontal opening to the environment; thus allows the consumer to see and handle the product that intends to acquire without inconvenience. The absence of this physical restriction for the consumer is accomplished by an air curtain that establishes an aerothermodynamics barrier between the conservation space, where the food products are located, and the surroundings. However, this merchandising solution contributes to the largest part of electrical energy consumption among refrigeration equipments installed in a typical supermarket [1]. The energy spent during the commercialisation of refrigerated food products is about 50% of the supermarket’s total energy consumption [2]. The world’s demand for commercial refrigeration equipment is projected to rise 4.6% per year until 2012 (value referred to 2008), being the sales market headed by reach-in and walk-in coolers, freezers and display cabinets [3]. The evaluation and analysis of the thermal

* Corresponding author. Tel.: þ351 275 329 759; fax: þ351 275 329 972. E-mail addresses: [email protected] (P.D. Gaspar), [email protected] (L.C. Carrilho Gonçalves), [email protected] (R.A. Pitarma). 1359-4311/$ e see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2010.11.020

performance of display cabinets in order to develop strategies and methods that promote its energy efficiency is necessary [4]. Consider the sketch of an open refrigerated display cabinet (ORDC) as shown in Fig. 1. The air was drawn by the fans located in front of the evaporator. The air passing through the evaporator is cooled below the conservation temperature of the perishable products exposed in the shelves of the equipment. This air mass flow rate is conducted to the rear duct, where part of it is discharged inside the conservation space at low velocity across the back panel perforation (PBP). The other part of this air mass flow rate will supply the air curtain, which develops vertically between discharge (DAG) and return (RAG) air grilles. The air curtain reduces the external infiltration of air from outside at higher dry temperature and specific humidity. The effectiveness of this aerothermodynamics barrier varies due to thermal and mass diffusive effects that affect the thermal entrainment. These effects depend on flow instabilities and boundary effects, among others, leading to a minor conservation quality of food products and greater energy consumption and costs since the ambient air infiltration load is around 72% of the cooling load of ORDC [5]. Several methods, experimental and numerical, are adopted by many researchers to evaluate the thermal performance of ORDC, and particularly of air curtain. Some researchers carried out Computational Fluid Dynamics (CFD) parametric studies based on two- and threedimensional models. Cortella et al. [6] and Navaz et al. [7] evaluated

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P.D. Gaspar et al. / Applied Thermal Engineering 31 (2011) 961e969

Nomenclature a1 to a5 b Cp g h hfg H Hc I L _ m Q_ R Re Ri T v W x, y, z X0 XPBP

Constants adopted in the correlation model [-] Air curtain width, [m] Specific heat, [J kg1 K1] Gravitational acceleration, [m s2] Enthalpy, [kJ kg1] vaporation enthalpy, [kJ kg1] Height, [m] Air curtain height, [m] Electrical current intensity, [A] Length, [m] Mass flow rate, [kg s1] Heat transfer rate, [kW] Central distance between DAG and RAG, [m] Reynolds number Richardson number Temperature, [K] Average velocity, [m s1] Width, [m]. Spatial coordinates (along length, width, height), [m] Thermal entrainment factor for air curtain without PBP airflow Thermal entrainment factor for perforated back panel airflow

Superscripts and subscripts amb Ambient cons Conservation zone DAG Discharge air grille

the influence of DAG velocity in thermal performance, quantifying the air infiltration through the frontal opening. Axell and Fahlén [8] developed a CFD parametric study to evaluate the influence on the thermal performance of air curtain height/width ratio and inlet velocity. Navaz et al. [9] calculated the amount of entrained air as a function of Reynolds number, based on jet width and velocity and inlet turbulence intensity, to evaluate the optimum operating conditions. Foster et al. [10] developed 3D CFD models to analyse the effect of changing the size and position of the evaporator coil, the width and angle of DAG and inserting baffle plates into the upper duct. D’Agaro et al. [11] carried out 2D and 3D CFD parametric studies to evaluate the influence of: longitudinal ambient air movement; display cabinet length, and air curtain temperature on the extremity effects and how it reflects in the ORDC performance. Chen [12] developed a CFD parametric study of length/width ratio and discharge angle of air curtains, height/depth ratio of the cavity and dimension and position of the inside shelves on thermal barrier performance of air curtains. Ge and Tassou [13] developed correlations for the heat transfer across air curtain with reasonable agreement with experimental data at steady state conditions, based on results obtained from a finite difference model. Other research works were just experimental, such as the study developed by Chen and Yuan [14] to evaluate the ambient air temperature and relative humidity; indoor airflow; DAG velocity; PBP airflow; and night-covers application, on the performance of an ORDC. Gray et al. [15] also conducted an experimental study to evaluate the effect of the perforation pattern of PBP on the distribution of airflow. Among the experimental techniques used by the researchers are thermocouple thermometry, hot-wire/film anemometry, laser Doppler anemometry, digital particle image velocimetry, hygrometry, tracer gases, and infrared thermography. Although these experimental techniques are reliable and provide

evap in lat m out PBP prod RAG sen surf

Evaporator Input; Upstream Latent Mixture Output; Downstream Perforated back panel Food products Return air grille Sensible Surface

Greek symbols Perforated back panel airflow ratio [e] Density, [kg m3] Dimensionless geometrical parameters [e] Relative humidity, [%] Direction of the ambient air velocity, [
] Dynamic viscosity, [kg m1 s1] Absolute humidity, [kgv kg1 a ]

b r p1, p2 4 q m u

Abbreviation 2D Two-dimensional 3D Thee-dimensional CFD Computational Fluid Dynamics DAG Discharge air grille ET Experimental Test ORDC Open refrigerated display cabinet PBP Perforated back panel RAG Return air grille TEF Thermal Entrainment Factor

a high degree of confidence in the CFD modelling approaches, its use involves a high cost and results are dependent on the ORDC geometry, DAG parameters and ambient air conditions. Other method that can be used to evaluate the thermal performance of an ORDC considering the thermal barrier provided by the air curtain is the thermal entrainment factor (TEF) calculation as proposed by [9,11,14] without PBP airflow or as proposed by Yu et al. [16] to consider the PBP airflow on calculation. This paper presents the analysis of experimental test results to evaluate the total sensible and latent heat transfers of an ORDC and the air curtain TEF for different environmental conditions. The main objective is to evaluate the influence of environmental conditions (air temperature, relative humidity and velocity, magnitude and direction) on the thermal performance of the ORDC. Also, it tries to clarify that TEF as it is defined cannot be used straightforward to assess the thermal entrainment for all air curtain applications. 2. Experimental testing 2.1. Experimental apparatus The ORDC experimentally tested is a self-contained system, in which the condensing unit and controls are built into the cabinet structure (beneath the cabinet, taking up the entire lower part). Its dimensions are 1900  796  1911 mm (L  W  H). It has four shelves and a well tray, being its frontal opening height, Hc, 1209 mm (see Fig. 1). The evaporator is housed under the well tray and integrated with refrigerant feed and return lines. It is an assembly composed by 18 tubes (16 mm of external diameter) and 135 fins/m. Its dimensions are 1600  360  130 mm (without tube curves), stretching over the entire length of the cabinet having an

P.D. Gaspar et al. / Applied Thermal Engineering 31 (2011) 961e969

Fig. 1. Configuration of ORDC and measuring probes locations.

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end clearance of about 100 mm aside. It is cooled by R404A refrigerant directly expanding in multiple passes. Temperature regulation and defrosting are digitally conducted by an electronic thermostat being the cooling system fan-assisted (four forward fans that blow air through the coil). The refrigeration cycle is controlled by the cold air temperature variation. This temperature is monitored by a limit positive temperature coefficient (PTC) thermistor probe located at half height of the rear duct. The cut-out temperature of the refrigeration system is set to 273 K. The thermostatic differential is set to 276 K, while the minimum and maximum values of setpoint temperature are set to 273 K and 277 K respectively. The evaporator is set to defrost at specific intervals (180 min). The controller provides direct compressor control through high power relays by stopping the coolant flow and starting the operation of an electrical resistance. Defrost period ends by time out (about 30 min maximum) or by reaching limit temperature setpoint of 285 K. This temperature measured by a defrost limit PTC thermistor probe located on the evaporator’s surface. Additionally, there are flow control devices, namely, a solenoid valve and an externally equalized thermostatic expansion valve, to modulate the cooling capacity delivered at the coil. The air temperature peaks occur each defrost cycle and influence the food temperature. The temperature increases during the refrigeration cycle (evaporator operation) due to the cooling capacity reduction with frost accumulation on the evaporator. Frost on the coil’s surface reduces velocity and increases temperature since the frost acts an insulator for heat transfer. During defrost cycle the refrigeration system does not work and the electrical resistance operates, which reflects into a large heat load that increases temperature. Furthermore, the thermal entrainment is much higher during this period since much more heat comes from outside warm environment. The frost accumulation was accounted by the condensate volume removed by the defrost system each cycle. Depending on the test conditions, the cooling load on the coil varies. This is due to RAG temperature and humidity values, but also to the frost formation pattern on the coil surface. The latter completely alters the patterns of flow and air temperature leaving the cooling coil. The experimental tests (ET) were performed in a climate chamber Aralab e Fitoclima 650000 EDTU. A Intab PC-Logger 3100 data acquisition system with the test probes described in Table 1 and placed inside the cabinet as shown in Fig. 1 was used to measure the physical properties such as temperature, relative humidity and velocity. A probe positioning system is used to evaluate the 3D effects of thermal entrainment on air curtain and properties variations along length and height of the conservation space [17]. The probe positioning system was settled in each shelf of the equipment and it

Table 1 Description of test probes and its location (see Fig. 1). Loc.

Type

Accuracy

Property

Ref.

Location

0e4

K-type thermocouple Hygrometer (odd measuring locations) K-type thermocouple Hot-wire anemometer Hygrometer K-type thermocouple Hot-wire anemometer Hygrometer K-type thermocouple K-type thermocouple (contact) K-type thermocouple (contact) Ammeter (Clamp-on)

0.5 K 3%

Temperature Relative humidity

Tcons

Conservation space

0.5 K 0.1 m s1(10%) 3% 0.5 K 0.1 m s1 (10%) 3% 0.5 K 0.5 K 0.5 K 0.02 A (2%)

Temperature Velocity Relative humidity Temperature Velocity Relative humidity Temperature Temp. (Surface) Temp. (Surface) Electric current

TDAG vDAG

DAG

TRAG vRAG

RAG

5

6

7 8 9 10 a

4cons

4DAG 4RAG

Tevap, Tevap, Tsurf I

out in

The surface temperature of the air sidewall pipe is measured at the evaporator inlet while at the outlet is measured the air temperature.

Evaporator outleta Evaporator inleta Interior surfaces Power source

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P.D. Gaspar et al. / Applied Thermal Engineering 31 (2011) 961e969

measured air temperature, relative humidity and velocity for three positions across the air curtain width and eight vertical crosssections along the equipment’s length. The positioning system moved the test probes in 240 mm increments for the 1800 mm length of shelves, taking 1 min to move between positions to reduce flow perturbation. Also, the properties values were acquire 1 min after reaching each position to ensure flow stabilization. The experimental results obtained with the point measuring technique along the air curtain and conservation space lengths show a similar behaviour to the experimental results obtained by [14,15,18]. It was found that the average variation in air temperature was 0.4 K and in air relative humidity was 4.5% which are not particularly significant. Based in these results, the probes were distributed in the mid plane of the equipment’s length. 2.2. Testing procedure The experimental testing follows the procedure defined by ENISO Standard 23953 [19], specifically in what concerns the Mpackage temperature class M1 (272.15K  Tprod  278.15 K). The temperatures were measured according to EN-ISO 23953-4 and 23953-5. The standard specifies test room climate classes, imposing air temperature and relative humidity as well as ambient air movement parallel (qamb ¼ 0
) to the frontal opening plane of the ORDC with a magnitude of vamb ¼ 0.2 m s1. The experimental tests (ET) were performed for some of these test room climate classes. Were also performed experimental tests with other environmental conditions (air temperature, Tamb, air relative humidity, 4amb, air velocity magnitude, vamb, and direction, qamb, relatively to the frontal opening of the ORDC). The environmental conditions of the experimental tests are described in Table 2. Before the tests were started, the cabinet was switched on and run during 24 h at the specified climate class (room air temperature, relative humidity and velocity) with the night-covers extended. Then, the cabinet was filled up to the load limits with test packages and M-packages. After this running-in period it was verified that stable operating conditions were reached as the air temperature curve agreed within 0.5 K, presenting no trend away from the mean temperature. This condition was verified for all tests, except for ET7 to ET9 that consider direction and magnitude of the room air velocity different from standard. The night-cover was removed and then the 24 h test period was initiated. Only the experimental data collected during the last 12 h of 24 h tests (initiated after the M-packages steady state temperatures were

accomplished) was considered. Table 2 contains the average mean values of the parameters measured during this period. 3. Heat transfer rate and thermal entrainment factor The heat transfer rate within an ORDC is given by the sum of the following individual sensible and latent thermal loads organized in a decreasing level of relevance: infiltration; radiation; internal sources as anti-sweat heaters, defrost system, lighting and fan motors; transmission (or conduction); and products subdivided into a pull-down load (due to their initial higher temperature before delivering the products into the cabinet and their temperature rise after a defrost cycle) and a products respiration load (for fresh fruits and vegetables) as exposed by [2]. However, Chen and Yuan [14] present simplified equations to calculate the total heat transfer rate (Eq. (1)):

Q_ T ¼ ðrvAÞDAG ðhRAG  hDAG Þ

(1)

This cooling load can be divided into its sensible (Eq. (2)) and latent (Eq. (3)) components. The sensible portion refers to temperature driven heat into the ORDC, whereas the latent portion refers to heat content of moisture entrained through air curtain and of product respiration:

Q_ Ts en ¼ ðrvAÞDAG Cp;m ðTRAG  TDAG Þ

(2)

Q_ Tl at ¼ ðrvAÞDAG hfg ðuRAG  uDAG Þ

(3)

Besides the heat transfer rate calculation, other method that can be used to evaluate the thermal performance of an ORDC, considering the thermal barrier provided by the air curtain, is the thermal entrainment factor (TEF) calculation as proposed by [9,11,14] defined in Eq. (4). The thermal entrainment factor will be zero if the air curtain has no entrainment: hRAG ¼ hDAG (unreachable condition) and it will increase with air enthalpy at RAG. When the return air is only from the surrounding ambient air, hRAG ¼ hamb, the thermal entrainment factor will reach unity. However, these studies considered that the specific heat capacity of air, Cp, is constant, so the average air enthalpy can be substituted by air temperature.

X0 ¼

hRAG  hDAG TRAG  TDAG z hamb  hDAG Tamb  TDAG

(4)

Table 2 Environmental conditions of the experimental tests and average mean values of the measured parameters. Experimental test Climate class (EN-ISO 23953) Location

1

2

3

Parameter

Unit

ET1

ET2

ET3

ET4

ET5

ET6

ET7

ET8

ET9

Tamb

[K] [%] [m s1] [o] [K] [%] [K] [%] [m s1] [K] [%] [m s1] [K] [K] [K] [A]

289.2 80.0 0.2 0 274.9 85.1 273.8 85.9 1.4 277.8 96.2 1.5 272.0 273.1 280.2 3.7

293.2 60.0 0.2 0 275.4 88.2 275.1 82.5 1.3 280.3 93.9 1.7 271.9 271.8 279.5 5.4

294.2 45.0 0.2 0 276.7 83.7 274.3 82.4 1.5 279.7 82.5 1.6 273.0 269.2 280.2 7.0

295.2 65.0 0.2 0 278.4 85.4 275.1 82.3 1.4 281.4 97.8 1.9 272.7 271.5 275.9 5.9

298.2 35.0 0.2 0 279.1 68.2 274.8 79.7 1.4 283.9 67.4 1.5 270.3 267.5 276.4 7.6

298.2 60.0 0.2 0 277.1 86.2 276.0 82.4 1.5 282.4 88.1 1.7 272.1 272.3 280.2 8.6

298.2 60.0 0.2 45 280.6 74.9 276.8 82.9 1.5 283.7 84.3 1.4 274.4 272.3 280.2 8.9

298.2 60.0 0.2 90 279.1 85.7 277.9 82.3 1.5 284.0 91.9 1.6 276.7 271.3 280.2 8.9

298.2 60.0 0.4 0 285.4 86.4 284.0 85.4 1.6 292.0 83.7 1.6 283.0 276.4 280.2 9.0

4amb vamb

qamb

0e4 1, 3 5

Tcons TDAG

6

vDAG TRAG

7 8 9 10

vRAG Tevap,out Tevap,in Tsurf I

4cons 4DAG 4RAG

P.D. Gaspar et al. / Applied Thermal Engineering 31 (2011) 961e969

Yu et al. [16] deducted the TEF formula with PBP airflow (Eq. (5)) based on TEF without PBP airflow.

TEF ¼ ð1  bÞX0 þ bX0 XPBP

(5)

Where, b is the PBP airflow ratio given by:

b¼

_ PBP m _ PBP þ m _ DAG m

(6)

And XPBP is the TEF for PBP airflow:

XPBP ¼

TPBP  TDAG Tamb  TDAG

(7)

These researchers stated that TEF for air curtain depends upon five factors groups being the last one less relevant: DAG supply (velocity and temperature); DAG geometry (width and height); PBP supply (airflow ratio and temperature); ambient air (temperature); and refrigerated load. Other factors such as: air curtain turbulence intensity; air curtain length; air curtain return nozzle width; ambient air relative humidity; and refrigerated load amount and temperature, are considered to possess a weak influence on TEF. Since the parameters needed to calculate PBP airflow ratio were not measured by the probe positioning system, the air mass flow rate through the PBP is determined by continuity equation (Eq. (8)):

_ PBP ¼ m _ RAG  m _ DAG m

(8)

The temperature of the air flowing through the PBP was evaluated by several manual measurements in different locations performed during the test period. The average mean PBP air temperature values obtained through manual measurements during the refrigeration cycle period for all experimental tests, were always 1.0 K lower than the temperature value at DAG with a standard deviation of 0.7 K and higher than the temperature value downstream the evaporator. So, the temperature value of the air flowing through the PBP is assumed to be 1.0 K below the DAG temperature value, while its relative humidity is assumed to be equal to the measured at DAG. Fig. 2 shows TEF calculation with and without considering the influence of the humid air composition. In average, TEF using temperatures (constant value of Cp), defined previously as X0, is approximately 16% lower than using enthalpies. The ambient air humidity is taken into account when enthalpies are used to calculate TEF. Thus, the influence of the humid air composition on thermal entrainment across the air curtain should be considered 1.0

965

when these analyses are performed. Hereinafter all TEF results use enthalpies. The results based on TEF formula with and without PBP are analysed. The original TEF without PBP airflow calculation (X0) was developed for air curtains usually settled in entrances of industrial, commercial or residential buildings to thermally separate spaces. ORDC use also the flow through the perforation of the rear duct in order to cool the products. The relationship between the amount of air let out from the perforations on the rear back panel of the cabinet and the air curtain, on the cabinet conservation temperature and RAG temperature is complex and not straightforward, depending on several parameters such as ambient air conditions, cabinet dimensions, particularly, DAG, RAG, PBP and height, and temperature, humidity and velocity of air at these locations. Fig. 3 shows the influence of the PBP air mass flow on the TEF calculation. If PBP airflow is not considered, TEF calculation (X0) provides always higher values (average of 21% higher). The results are consistent with the physical phenomenon of “plug flow” provided by PBP that helps to stabilize the air curtain and carries heat on the conservation zone towards it. So, it is important that the PBP airflow is taken into account, when TEF is used to analyse the thermal performance of ORDC. 4. Results and discussion 4.1. Average contribution of individual thermal loads Table 2 contains the average mean values of measured parameters. The average air temperature and relative humidity inside the conservation space were obtained averaging the data collected by the probes located in position ner 0 to ner 4 (see Fig. 1). Likewise, the average temperatures of inner surfaces of the ORDC were obtained by averaging the values measured by probes ner 9 on upper, PBP and well tray surfaces. Using the equations proposed by Faramarzi et al. [20], the individual thermal load components are calculated for each experimental test. The contribution range of each component is presented in Fig. 4. The average contribution of infiltration of warm and moist ambient air across air curtain is about 73% of the cooling load for the experimental tests performed. So, it is clear that the barrier ability of the air curtain plays a significant role in thermal interaction of an ORDC and surrounding environmental conditions. These results are in accordance with test results obtained by [2].

1.0

+20%

0.8

+20%

0.8 -20%

0.6

0.6 ET1 ET2 ET3 ET4 ET5 ET6

0.4

0.2

0.0 0.0

0. 2

0.4

0.6

0. 8

1.0

TEF(T )

X0

TEF(h )

-20%

ET1 ET2 ET3 ET4 ET5 ET6

0.4 0.2 0.0 0.0

0.2

0.4

0.6

0.8

1.0

TEF Fig. 2. Comparison of TEF calculation with and without considering the influence of the humid air composition.

Fig. 3. Comparison of TEF calculation with PBP and without PBP (X0).

P.D. Gaspar et al. / Applied Thermal Engineering 31 (2011) 961e969

Fan motors 6% - 9%

Lighting 7% - 10%

Transmission 2% - 4%

Radiation 6% - 12%

Infiltration 67% - 77%

Heat transfer rate components [%]

966

100% 90% 80%

12%

70% 50% 40%

77%

67%

30% 20% 10% 0%

4%

3%

ET5

ET6 Experimental test

Fig. 6. Heat transfer rate components comparison for ( ) ET5 [4amb ¼ 35%] and (-) ET6 [4amb ¼ 60%] (Tamb ¼ 298.15 K; vamb ¼ 0.2 m s1; qamb ¼ 0
).

4.2. Total heat transfer rate and thermal entrainment factor The comparison between sensible and latent components of total heat transfer rate of the ORDC for each experimental test is presented in Fig. 5. The components were calculated making use of Eqs. (2) and (3). 4.2.1. Influence of ambient air relative humidity on heat transfer According to Faramarzi et al. [20], there is a direct relation between indoor relative humidity and mass of the moisture removed from air during the refrigeration process. This conclusion is easily identified comparing the proportion of sensible and latent heat transfer rates for experimental tests ET5 and ET6, that were performed at the same temperature (Tamb ¼ 298.15 K) but with different relative humidities, 4amb, 35% and 60% respectively. The latent heat transfer rate component increases 42%. The infiltration heat transfer rate is the individual component with higher increase. The heat transfer rate components comparison is shown in Fig. 6. The energy consumption, total heat transfer rate and ice formation on the evaporator coil surface increase with latent heat transfer rate. On environmental conditions nearest to saturation point, the refrigeration system extracts more mass of the moisture from air. With the relative humidity increase from 35% to 60% at constant temperature, the condensate volume removed by the defrost system each cycle increase from 0.8 dm3 to 2.5 dm3. The electrical power is not only consumed by the compressor but also includes the electrical resistance operation. ET6 has a higher latent load fraction than ET5, i.e., there is more frost accumulation on the coil surface, thus the electrical resistance operated during an extended time period to extract the frost.

4.2.2. Influence of ambient air temperature on heat transfer The value of total sensible heat transfer rate modifies with environmental conditions, but its variation is not so significant since it depends only on conservation zone and ambient air temperatures. Comparing the experimental tests ET2 and ET6 (see also Fig. 5), an increase of 5 K of ambient air temperature, Tamb, and consequently different temperature, relative humidity and velocity values at DAG, RAG and evaporator air off, determines an increase of the infiltration load of 30% (300 W). DAG velocity value is different between tests not only due to measurement errors (accuracy and averaging method). It must be taken into account that this cabinet is designed to provide the best performance at the ambient air conditions provided by test ET6, such as compressor operation, controlled amount of frost accumulation on the coil surface, time of defrost system operation. This fact ensures a higher DAG velocity for this environmental condition. Since TEF calculation does not take into account velocities, the proportionality relation between enthalpies at DAG, RAG, PBP and ambient environment will provide the same value (TEF ¼ 0.25). 4.2.3. Influence of ambient air direction on heat transfer Using the probe positioning system [17], the TEF with PBP is calculated along the non-dimensional length, x/L, of the ORDC. For the climate class ner 3 (Tamb ¼ 298.15 K, 4amb ¼ 60%, vamb ¼ 0.2 m s1 and qamb ¼ 0
), the ambient air movement from left to right in Fig. 7 likewise Fig. 8a, will increasingly promote thermal entrainment along length. The high values of thermal entrainment at the nearest sidewall locations (x/L ¼ 0.03 and x/L ¼ 0.97) can be attributed to 1.0

3.0 Sensible heat [kW]

0.9

Latent heat [kW]

2.5

0.8 1.2

0.7

2.0

0.6

1.5

0.8

1.0

1.0

TEF

Heat transfer rate [kW]

Fan motors Lighting Radiation Infiltration Transmission

60%

Fig. 4. Contribution range of each individual thermal load component to the refrigeration load of the ORDC depending on the tests environmental condition.

0.9

1.2

1.0

6% 7% 7%

8% 9%

0.7

0.4

0.8

0.6

0.5

1.7 1.0

0.6

0.8

0.9

1.0

0.3 0.2

1.1

0.7

0.6

0.5

0.1 0.0

0.0 ET1

ET2

ET3

ET4

ET5

ET6

ET7

ET8

ET9

Experimental test Fig. 5. Sensible (-) and latent ( ) total heat transfer rate [kW] comparison for the experimental tests.

0.03

0.17

0.30

0.43

0.57

0.70

0.83

0.97

Non-dimensional length, x /L Fig. 7. Thermal entrainment factor variation along non dimensional length (x/L) for climate class ner 3 (Tamb ¼ 298.15 K, 4amb ¼ 60%, vamb ¼ 0.2 m s1 and qamb ¼ 0
).

P.D. Gaspar et al. / Applied Thermal Engineering 31 (2011) 961e969

967

Fig. 8. Experimental test setups with different ambient air direction. (a) ET6 (qamb ¼ 0
); (b) ET7 (qamb ¼ 45
); (c) ET8 (qamb ¼ 90
).

sidewall effects, as the air curtain is unable to restrict the free entry of external air at the extremities of the ORDC. This is corroborated by experimental results [11]. The infiltration heat transfer rate is dependent on the direction of ambient air velocity, qamb. Fig. 8 shows the layout schematics for the ambient air direction on the experimental tests: ET6 with parallel direction (qamb ¼ 0
), ET7 with oblique direction (qamb ¼ 45
) and ET8 with perpendicular direction (qamb ¼ 90
) to the ORDC frontal opening. The values of ambient air temperature, relative humidity and velocity were maintained constant (Tamb ¼ 298.15 K, 4amb ¼ 60% and vamb ¼ 0.2 m s1). The total heat transfer rate increases 5% when the air movement direction goes from parallel to oblique and increases an additional 1% when it goes from oblique to perpendicular (see also Fig. 5). Considering the experimental data shown in Table 2 and the heat transfer rate results in Fig. 5, it is shown that the air curtain break due to direction of ambient air movement, either oblique or perpendicular to the ORDC frontal opening, promotes the thermal interaction between the conservation zone and ambient air masses. This condition reflects on an increase of air temperature between shelves. The results suggest that both oblique and perpendicular directions of ambient air lead to a similar increase of the heat transfer rate (approximately 100 W). The electrical energy consumption of the equipment increases about 4% due to increase of the refrigeration system operating time. The longer operation time of the refrigeration system increases condensate volume from 2.5 dm3 to 6.0 dm3 when the ambient air direction moves from parallel (ET6) to perpendicular (ET8). The TEF results for these experimental tests are shown in Fig. 9. These results are in accordance with results of [11,17,21]. From the

1.0 0.9 0.8 TEF

0.7 0.6 0.5 0.4 0.3

0.32

0.25

0.30

0.2 0.1 0.0 ET6 ( 0º )

ET7 ( 45º )

ET8 ( 90º )

Direction of the ambient air velocity, θ amb [ º ] Fig. 9. Comparison of TEF calculation for the direction of the ambient air velocity parallel (ET6: qamb ¼ 0
), oblique (ET7: qamb ¼ 45
) and perpendicular (ET8: qamb ¼ 90
) with Tamb ¼ 298.15 K, 4amb ¼ 60% and vamb ¼ 0.2 m s1.

viewpoint of physics, it is easier for ET8 to affect the sealing ability of air curtain. Therefore, it was expectable a larger TEF for ET8 than that for ET7. However, due to the TEF formulation based on a balance between the enthalpies difference at external environment (ambient air) and outflow (RAG and PBP) relatively to inflow (DAG), the TEF result is smaller for ET8 (0.30) than for ET7 (0.32). The recirculated air curtain of the ORDC contributes to these results, since the PBP air (low temperature) at lower height of the air curtain is added directly to the mass flow that enters RAG, lowering the air temperature at this location. This situation is minor for ET8, being the thermal entrainment higher than for ET7 only at the middle height of the opening. Furthermore, there is a superposition of physical effects beyond the resultant from the direction of ambient air. So, the TEF calculation led to obtain results in contradiction with the physical perspective. This is verified by the analysis of heat transfer, electrical energy consumption, condensate volume and ORDC air temperature higher values for perpendicular (ET8) than for oblique (ET7) ambient air direction to the ORDC frontal opening. Again, the comparison of these TEF results shows that as it is defined cannot be used straightforward to assess the thermal entrainment for all air curtain applications. 4.2.4. Influence of ambient air velocity magnitude on heat transfer This section considers the analysis of heat transfer rate values for the experimental tests when the magnitude of ambient air velocity is increased from vamb ¼ 0.2 m s1 (ET6) to vamb ¼ 0.4 m s1 (ET9), maintaining the values of ambient air temperature, relative humidity and direction of velocity constants (Tamb ¼ 298.15 K, 4amb ¼ 60% and qamb ¼ 0
). ET9 intends to simulate ambient air conditions that can be found in small grocery stores with reduced dimensions and sometimes without an air conditioning system. The door opening promotes pressure differences that can originate air currents. Test ET9 simulates an air current or the wind entrance through the front door, characterized by an air velocity magnitude increase, whose direction is parallel to the frontal opening plane of the display cabinet. As shown in Fig. 5, the increase of ambient air velocity magnitude, even parallel to the plane of the equipment’s frontal opening, promotes thermal interaction between the conservation zone and ambient air masses by disturbance of aerothermodynamics barrier provided by air curtain. The total heat transfer rate increases 53% due to increase of air infiltration load across the air curtain. This result is verified by nearly three times increase of TEF as shown in Fig. 10. The experimental results described here and the numerical results of [22,23] show that the equipment end walls promote air curtain instability and consequently thermal entrainment. The air curtain instabilities are more significant from the side where the ambient air comes in. There is a significant cold air leakage near the lateral wall of the incoming ambient airside because it causes both intense turbulence and

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1.0 0.9 0.8 0.64

TEF

0.7 0.6 0.5 0.4 0.3

0.25

0.2 0.1 0.0 ET6 ( 0.2 ms-1)

ET9 ( 0.4 ms-1)

Ambient air velocity, v amb [ms-1] Fig. 10. Comparison of TEF calculation for different magnitudes of the ambient air velocity (ET6: vamb ¼ 0.2 m s1), (ET9: vamb ¼ 0.4 m s1) with Tamb ¼ 298.15 K, 4amb ¼ 60% and qamb ¼ 0
.

vortex formation. The vortices formation occurs after the ambient air overcomes the equipment’s lateral wall, increasing the thermal entrainment mainly near the equipment’s extremities. Additionally, the air curtain stability decreases with distance to the DAG. Eddies are developed by shear layer interactions and enhanced by turbulence intensity at the initial region of the air curtain jet that triggers mixture. Practically no air curtain exists near the RAG since its momentum has been reduced along the way. The mixture between the perpendicular air curtain and ambient air masses also promotes vortices that increase the thermal entrainment near the end walls and also across the air curtain. So, there is a reduction of the air curtain aerothermodynamics performance as the ambient air velocity increases, being the thermal entrainment downward the air curtain dependent of the eddies formation. These conditions lead to a non-uniform distribution as well as an increase of air temperature in the conservation zone above the higher limit for the proper conservation of food products. The higher mixture between refrigerated air and warm and moist air of external surroundings, increases the ice formation on the evaporator coil surface, and consequently the condensate volume (from 2.5 dm3 to 9.0 dm3). This condition also leads to an increase of electrical energy consumption in approximately 5%. Additionally, the ambient air movement affects the return air temperature and consequently the energy efficiency of the equipment. The influence of increasing the magnitude of ambient air velocity or its direction relatively to the frontal opening of the equipment reflects in an improper performance of the ORDC, being unable to maintain the value of air temperature at the conservation zone below the acceptable limit to ensure food safety. Considering that the tested ORDC has as technical characteristics a heat extraction rate of almost 2.5 kW (compare this value with the results shown in Fig. 5) for a refrigerant (R404A) evaporating temperature of 263.15 K and refrigerant condensing temperature of 318.15 K, the increase of heat transfer individual constituents, mainly ambient air infiltration thermal load (and consequently the increase of TEF), due to modifications of ambient air velocity direction or its magnitude cannot be suppressed by the refrigeration system. The variation of product temperature is also a good indication of load on the ORDC. The final product temperature exceeded largely 278.15 K in ET7 to ET9. The product temperature increased in average, per refrigeration cycle, 0.6 K, 0.9 K and 1.8 K respectively until stabilizes at constant values. Although, the variation of product simulator temperature within the ISO class 3 condition (ET6) is similar to the experimental results obtained by [10,15,18,24,25].

4.2.5. Final remarks The analysis of experimental results based on heat transfer rate and thermal entrainment factor calculations allow drawing some relevant conclusions of heat and mass exchange process between environmental and conservation zone air masses: (1) increase of relative humidity causes an immediate impact in latent heat load. The thermal entrainment factor should be calculated using enthalpies rather than temperatures to include the humidity influence; (2) air temperature variation influences the sensible heat load; (3) air velocity magnitude increase, with direction parallel to the frontal opening plane, promotes thermal and mass interaction due to disturbance of the aerothermodynamics system of air curtain, resulting in an increase of total heat transfer and of thermal entrainment factor; (4) changes of ambient air velocity direction from parallel to perpendicular provides a significant increase of latent and sensible heat transfers. Tests performed beyond standard conditions (ET7 to ET9), contain physical phenomena related with air mass entrance across the frontal opening at its middle height that will increase the thermal entrainment. However, the refrigerated air mass entering the display cabinet through the perforated back panel at lower height (higher velocity and lower temperature) is directed to the RAG performing a “short-circuit”, that reduces the temperature value at this location and lead to higher TEF values. In this kind of tests, the use of TEF to evaluate thermal entrainment must be very cautious. 5. Conclusions This experimental work was conducted to study the heat transfer rate and thermal entrainment factor of air curtains in vertical open display cabinets for different ambient air conditions, varying air temperature, relative humidity and velocity (magnitude and direction). The analysis of experimental results based on heat transfer rate and thermal entrainment calculations allow drawing some relevant conclusions of heat and mass exchange process between environmental and conservation zone air masses. The thermal entrainment factor (TEF) calculation is an engineering approximation to evaluate the thermal performance of a display cabinet instead of calculating all sensible and latent thermal loads. Besides this advantage, it is a dimensionless number, providing a way to compare thermal performance between appliances. However, considering the results provided in this paper, TEF cannot be used indiscriminately because it only takes into account the enthalpies difference between inflow (DAG) and outflow (RAG). From an engineering point of view, TEF can be used to design a better cabinet through its calculation while testing fans with different blade configuration and dimensions as well as different velocity, under the same climate class condition. These experimental tests should look for the optimum DAG velocity that ensures the minimum thermal entrainment. In conclusion, this study shows a strong dependence of thermal and mass interaction processes in open refrigerated display cabinets to variations of surroundings ambient air conditions. Therefore, it is essential that the conception and design of these equipments take into account technical solutions to provide an optimum discharge air velocity as well as to reduce the vortices formation near the lateral walls and moving them away from the air curtain. Design modifications in the lateral walls, display cabinet structure and devices aim to reduce the mixture of ambient air with refrigerated air along and across air curtain and in sidewalls. Additionally, the store HVAC system where the fixture will be installed must take care of the possible variation of ambient air parameters. These conditions must be followed by a wise and prudent positioning of the open refrigeration cabinet relatively to

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