recent developments in vapor compression

Proceedings of the ASME 2011 9th International Conference on Nanochannels, Microchannels and Minichannels ICNMM2011 June 19-22, 2011, Edmonton, Alberta, CANADA

ICNMM2011-58277 RECENT DEVELOPMENTS IN VAPOR COMPRESSION TECHNOLOGIES FOR SMALL SCALE REFRIGERATION APPLICATIONS Jader R. Barbosa, Jr. Polo – Research Laboratories for Emerging Technologies in Cooling and Thermophysics Department of Mechanical Engineering Federal University of Santa Catarina (UFSC) Florianópolis, SC, Brazil [email protected]

ABSTRACT This paper reviews the current trends and research efforts associated with the development of miniaturized mechanical vapor compression refrigeration systems. Over the past decade, there has been a significant number of studies devoted to the miniaturization of its components, the most critical being the compressor. The paper will focus on the thermodynamic and thermal aspects of the development of small compressors and other components of small-scale cooling cycles. Whenever appropriate, issues and challenges associated with different cycle and component designs will be addressed. INTRODUCTION For near room temperature applications, mechanical vapor compression refrigeration exhibits the highest values of the coefficient of performance (COP) among the existing refrigeration technologies (e.g., solid-state, sorption, gas cycles etc.). The COP of a refrigerator is defined as the ratio of the desired heat output rate from the low temperature source to the required rate of work input to the compressor. Although mechanical vapor compression has widespread use in several industrial sectors, only in the past decade or so, due to improvements in the synthesis of new materials and fabrication techniques, it has become possible to envisage its application in refrigeration of small-scale systems. Applications of small-scale cooling systems are vast and range from automotive to medical, military to aerospace, and personal to electronics cooling. In this review, due to space limitations, we will focus on recent developments associated with electronics cooling. As far as the refrigeration system is concerned, although some aspects of the cycle and of other components are addressed, the review is centered on the

compressor. A more thorough review that includes a wider range of related subjects is presented in [1]. In cooling of electronics, low temperature and reliability are synonymous. In addition to making the switching times of semiconductor devices faster, the temperature reduction can increase the circuit speed due to the lower electrical resistance of the interconnecting materials. It can also reduce the thermally induced failures of devices and components [2]. Controlling the junction temperature is also crucial for reducing leakage currents, which increase exponentially with temperature. According to the analysis of Copeland and Chan [3], at the usual junction temperatures of near 85oC, leakage currents were expected to become as much as one-third of the total power dissipation. Additionally, attainable performance improvements range from 1 to 3% for every 10oC reduction of the transistor temperature, depending on the chip characteristics [2]. Ways of reducing the thermal resistances in passive1 cooling systems have improved substantially in recent years, and its evolution has been reviewed in a number of articles [48]. However, with the decreasing length scale of electronic equipment, reducing the thermal resistances will eventually reach an asymptotic limit [9]. As pointed out by a number of researchers [10-13], natural and forced convection of air are no longer capable of maintaining processor temperatures below acceptable values. Other passive cooling technologies will also eventually reach their maximum cooling capability. There is an understanding that mechanical vapor compression

1 The word passive is used here to describe a cooling technology that does not rely on refrigeration. Thus, in principle, an active cooling technology is capable of lowering the chip temperature below the ambient temperature.

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Copyright © 2011 by ASME

refrigeration is among the most promising active cooling technologies to replace conventional heat sinks. In an active cooling system, a net work input is used for lifting the thermal load applied on the low temperature source. The work input plus the thermal load are rejected as heat into the ambient. A typical single-stage vapor compression-based processor cooling system is composed of a compressor, a condenser, an expansion device and a heat sink/evaporator, as seen in Fig. 1.

R j amb 

* T j  Tamb

q

(1)

can become negative if the junction temperature is also lower than the ambient. The rate a which heat is rejected at the condenser is given by: T T* (2) qc  q  w  cond amb Rc  amb where the cooling capacity can be written in terms of the heat sink/evaporator thermal resistance, Rj-evap, as follows: q

(a)

(b)

Figure 1. (a) Schematic diagram of a vapor compression cooling system for microprocessors. (b) Thermal resistance network. Rj-evap encompasses the thermal resistances between the chip and the evaporating fluid. Depending on the operating conditions, the volatile refrigerant in the heat sink can boil at a temperature lower than that of the ambient. Thus, the equivalent junction-toambient thermal resistance:

T j  Tevap R j  evap

(3)

In the above equations, qc is the heat rejection rate at the condenser, w is the rate of work consumption at the compressor, Tevap is the evaporating temperature of the refrigerant, Tcond is the condensing refrigerant temperature, * is the effective ambient temperature (i.e., corrected for Tamb effects of proximity and pre-heating) and Rc-amb is the thermal resistance between the condensing refrigerant and the ambient. According to Phelan et al. [14], the potential advantages of refrigeration cooling, in general, are the following: (i) ability to dissipate heat while maintaining a low junction temperature; (ii) higher device speed and increased reliability because of the reduced operating temperature; (iii) expanded device lifetime because of the constant operating temperature. On the other hand, there are several issues that need to be addressed before a successful implementation of vapor compressor technology for electronics cooling in particular. These are as follows: (i) designing an efficient, reliable and compact compressor, (ii) developing and improving suitable cold plate evaporators, (iii) integrating the cooling system in a restricted space, (iv) resolving packaging related issues, and (v) maintaining a competitive cost for the entire system. In contrast with other passive cooling techniques (such as heat pipes and single-phase liquid loops), sprays, two-phase jets and boiling in microchannels can, in principle, be easily combined with vapor compression cooling in order to achieve junction temperatures below the ambient temperature. As can be seen in this and other reviews [1] with respect to the evaporator geometry to be used in small-scale refrigerators, there seems to be a consensus in the literature that evaporator designs with small parallel channels are capable of providing the desired cooling capacity for such applications. Nevertheless, the performance of evaporators and heat sinks can be improved by a more fundamental understanding of the boiling processes in small channels coupled with novel design methods and advanced fabrication techniques.

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As far as the compressor is concerned, a number of different designs based on distinct working principles have been proposed for application in small systems. This review will point out their main characteristics and will present future directions of their research. NOMENCLATURE

air-cooled microchannel condenser. The evaporating temperatures ranged from 5 to 20oC, and the desired heat spreader temperature was also 20oC. The COP ranged from 2.2 to 5.8. The compressor was driven by a brushless 24 V DC motor. The swept volume was 1.8 cm3 and the compressor speed ranged from 3000 to 4000 rpm. The compressor external dimensions were 8.9 cm (dia.) and 6.4 cm (height).

Roman COP Coefficient of performance [-] q Evaporator/heat sink heat transfer rate [W] qc Condenser heat transfer rate [W] R Thermal resistance [oC W-1] T Temperature [oC, K] w Compressor power [W] Greek η Efficiency [-] Subscripts amb Ambient C Carnot cond Condenser/condensing refrigerant evap Evaporator/evaporating refrigerant E External I Internal j Chip/junction/copper block/chilled water R Refrigerator

(a)

SURVEY OF MECHANICAL VAPOR COMPRESSION SYSTEMS FOR COOLING OF ELECTRONICS Review of Refrigeration Systems This section reviews the experimental studies on the development of working systems and prototypes for cooling of electronics using mechanical vapor compression technologies. A more thorough assessment of the literature, including modeling studies and other applications (e.g., personal cooling), is available in [1]. Schmidt and Notohardjono [15] designed and tested a modular vapor compression refrigeration unit to cool the CMOS processors of a computer server. The system employed a sliding vane rotary compressor with a brushless DC-motor and a thermostatic expansion valve. The overall system volume was approximately 51800 cm3, and it provided cooling capacities in the range of 650 to 1050 W for average evaporating (cold plate) temperatures between 15 and 35 oC, using R-134a. Reinforced materials were used inside the compressor to provide a longer life for the system. Maveety et al. [16] constructed a small-scale refrigeration system for a 2-U (1-U = 44.45 mm or 1.75 in) rack that used R-134a and was capable of providing cooling capacities of up to 130W (see Fig. 2). The system was composed of a copper cold plate, a small rotary compressor, a capillary tube and an

(b)

Figure 2. (a) The system of Maveety et al. [16]. (b) Behavior of COP as a function of Tevap and Tamb. Bash et al. [17] constructed a microprocessor refrigerator prototype based on a variable capacity acoustic compressor. The system was tested with a model 5-U server with four heater blocks capable of delivering up to 100 W of thermal load each. The power dissipated in each block was controlled independently. The oil-less, 780 Hz, acoustic compressor operated with R-134a. The four serpentine evaporators were connected in series and were attached to each heater block. The system exhibited controllability of the evaporating temperature of the four heaters to within 2oC of the set temperature under uniform heating, for thermal loads between 20 and 80 W per processor. The acoustic compressor (see Fig. 3) is based on the resonant macrosonic synthesis technology that allows the creation of standing waves in an acoustic resonator whose shape is responsible for the formation of high-amplitude

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shock-free waves. The excitation amplitude of the variable reluctance linear driver controls the pressure ratio and the capacity of the compressor. The compressor is equipped with automatic reed valves for refrigerant suction and discharge.

refrigeration efficiency of the system (to be defined later) varied between 33% and 52%. They used a commercially available rotary compressor (R-134a) with variable speed in the range of 2000 to 3500 rpm. The compressor swept volume was approximately 2.3 cm3. Failure of the compressor has been reported after 400 h of operation, which adds force to the fact that reliability and efficiency of the compressor are the main challenges concerning the miniature-scale compressor design.

Figure 3. Schematic diagram of the acoustic compressor [17]. Mongia et al. [9] developed a R-600a vapor compression refrigeration system for portable computers. The compressor was a positive displacement (linear motor) 12 V DC miniature model with a volume displacement of 12.5 cm 3/s. Microchannel heat exchangers (evaporator, superheater and condenser) were employed (see Fig. 4). The system was capable of lifting thermal loads of up to 50 W, 80% of which have been transferred in the evaporator. A coefficient of performance of 2.25 was obtained for evaporation and condensation temperatures of 50°C and 90°C, respectively. The calculated internal refrigeration efficiency of the system (defined as the ratio of the actual COP and that of a Carnot refrigerator operating between the same condensing and evaporating temperatures) corresponded to 25-30% for the approximately 15 experimental tests. It is worth noting that external air cooling was needed during the tests to keep the compressor case temperature below 70oC. Compressor isentropic efficiencies in the range of 0.25 to 0.4 were reported for values of pressure ratio between 1.5 and 2.7, approximately. Trutassanawin et al. [12] carried out an experimental investigation of a miniature vapor compression refrigeration system. The evaporator consisted of 41 parallel channels with a cross-section area of 0.8 mm x 2.3 mm, while the condenser consisted of 20 parallel channels with a cross-section area of 0.62 mm x 0.33 mm. The following operating conditions were evaluated: evaporation temperature of 10 to 20°C, compressor inlet refrigerant superheating of 3 to 8°C, condensation temperature of 40 to 60°C and condenser outlet subcooling of 3 to 10°C. The experimental results showed that the internal

Figure 4. Test system used by Mongia et al. [9]. Agwu Nnanna [18] investigated the transient response of a vapor compression cooling system to rapid changes in thermal load typical of high heat flux electronics and high-end computers. A condensing unit equipped with a scroll compressor was used. The evaporator consisted of a four-pass 8 mm ID, 10 mm OD coil mounted on a 152.4 x 88.9 x 19.1 mm aluminium block which could be operated under uniform and non-uniform heating conditions. An internally equalized thermostatic expansion valve was selected as the expansion device. The cooling capacities evaluated in the experiment ranged between 152 and 606 W. The vapor compression system was capable of maintaining the junction temperature of the simulated electronics below the value usually attained by conventional air-cooling systems. The authors indicated large spatial temperature gradients on the evaporator surface as a result of outlet refrigerant superheating. Also, they reported on temperature oscillations at the evaporator and thermostatic expansion during the early stage of the cooling process. Rao et al. [19] described a series of experiments to demonstrate the feasibility of a miniature-scale refrigeration system for cooling of a four-chip (quad) CPU. The system consisted of a 70 W R-12 compressor (possibly reciprocating – no specific information is given in the paper) with an aircooled condenser and a capillary tube. Four aluminium heater blocks were used to emulate the microprocessor heat sinks. Four evaporators connected in series were attached to the heater blocks. No information on the geometry of the

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evaporators has been provided. The system was capable of lifting thermal loads of 120, 150 and 180 W, while maintaining the temperatures of the four heaters as low as 20°C for the 150-W case. However, the time taken for the system to reach steady-state with a uniform distribution of temperature among the heaters was of the order of one hour. The reported values of COP ranged from 0.32 to 0.58 for the cooling capacities of 90 and 180 W, respectively. The authors reported that the major refinements of the system are the fan noise and the size of the compressor. Coggins [20] and Coggins et al. [21], building upon previous work due to Wadell [22], designed a two-stage cascade refrigeration system (with R-508b and R-404a) for cooling of high performance microprocessors. The evaporator was manufactured from copper with 13 parallel microchannels of 600 μm hydraulic diameter. The lowest evaporating temperature reported by Coggins [20] was -82oC (no load), while the highest cooling capacity was 40 W. Under this condition, the reported evaporating and chip surface temperatures were -72.3 and -60.5°C. The authors pointed out the need for smaller and more powerful compressors in order to develop more compact and more silent systems. Review of Compressors In addition to the information already presented in the previous section, this section provides further information regarding compressors that have been modeled, designed, constructed and/or tested for small-scale applications. It also reviews studies that were either solely dedicated to the description of a particular compressor or whose level of detail in the description of the cooling system was insufficient or too speculative. The compressor is certainly the most challenging component of the refrigeration system in terms of its miniaturization. Studies that have dealt with applications in electronics cooling but were not too focused on the actual system miniaturization, used small (but not truly miniaturized) commercially available compressors. Since some commercial compressors are designed for operating at large pressure ratios (e.g., refrigerators and other low back-pressure systems), their performance is poor at the relatively small pressure ratios associated with the high back-pressure conditions often desired for electronics cooling. As far as the compression technologies for the miniaturized compressors are concerned, a clear-cut decision has not yet been reached as to which working principle (rotary, reciprocating, linear etc.) is best for the several types of specific applications in electronics and/or personal cooling. Regarding the power supply, if an electric motor is used, operating with direct current minimizes the start-up power in comparison with alternating current. The peak current can be several times higher than the operating current if an AC single-phase motor is used [15].

Rotary compressors have been used by a number of authors [12,15,16]. More recently, a rolling piston variable speed compressor (2000 to 6500 rpm) was investigated experimentally in a hot-gas cycle calorimeter by Sathe et al. [23]. In this apparatus, no change of phase takes place. The suction pressure and temperature are controlled by valves and the discharge pressure is controlled by the refrigerant charge. The tests were carried out for evaporating temperatures between -18 and 24oC, and condensing temperatures between 27 and 71oC. The compressor has reached volumetric efficiencies in the range of 73% to 90% and insentropic efficiencies between 44% and 70%. The cooling capacity was rated from 163 W to 489 W, while the COP varied from 2.1 to 7.4. The isentropic and volumetric efficiencies were not influenced significantly by the compressor speed. On the other hand, the theoretical cooling capacity showed a strong linear dependence on the compressor speed. Sathe et al. concluded that, due to the improved performance in comparison with two oversized compressors for small-scale applications [12], the small rotary compressor had the potential for use in miniature vapor compressor refrigeration systems. A great deal of effort has also been directed at the development of linear compressors for small applications such as electronics cooling [24-28]. Amongst the advantages of linear driving mechanisms, one can cite (i) reduced frictional losses in comparison to conventional reciprocating compressors (due to the absence of conversion of rotating motion into linear motion), (ii) oil-free (or very low viscosity oil) operation, and (iii) smaller motors due to operation at the resonant frequency [26]. The absence of lubricant oil enables gravity-independent operation, which is an important feature for mobile electronics. Unger and Novotny [24] presented the design of a linear compressor in an opposing configuration (Fig. 5) for a 2U-rack mount. The following operating conditions were considered: evaporating and condensing temperatures of 20 and 60 oC, cooling capacity of 1500 W and ambient temperature of 45 oC. R-134a was used as the refrigerant. The compressor was driven by permanent magnet motors and an oscillating current coil. The springs of the planar type and lubrication was achieved by gas bearings. The suction valves were of the inertia-type mounted on the face of the opposed pistons. The discharge valves were mounted radially from the workspace through the cylinder wall. The compressor was tested in a hot-gas cycle calorimeter, and its COP, calculated as the product of the measured mass flow rate and the enthalpy of vaporization at the suction pressure is shown in Fig. 6. Wang and Tai [25] presented a prototype of a linear compressor driven by a magnetic actuator (see Fig. 7). Ideally, the compressor was designed to operate at a resonant frequency of 56 Hz, with a corresponding piston stroke of 5.8 mm. However, due to friction, clearance and leakage losses,

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the prototype was only capable of delivering only a fraction of its intended performance.

Coil

Piston

Magnet Spring

Discharge

Suction Valve

Figure 5. The linear compressor of Unger and Novotny [24].

prototype was tested in a hot-gas cycle calorimeter with R134a for suction and discharge pressures ranging from 534/561 to 587/709 kPa. The frequency was 44 Hz and the pumped mass flow rate ranged from 0.12 to 0.46 g/s. The agreement between the model and the experimental data was somewhat above ±20% (the isentropic and volumetric efficiencies ranged from 2.8 to 5.7% and from 14.1 to 50.5%, respectively), and was reported to be sensitive to small changes in the values of parameters that are difficult to quantify (e.g., size of the leakage gap, eccentricity etc.). The Brazilian company Embraco [28,29] has recently developed two miniaturized orientation-free linear compressors for application at high back-pressure (HBP) and medium backpressure (MBP) conditions (see Fig. 8 and Table 1 for dimensions). The cooling capacity and COP curves for MBP conditions (using R-134a) obtained in a hot-gas cycle calorimeter built at the Federal University of Santa Catarina are shown in Fig. 9. The operating frequency at Tevap/Tcond of 15/55oC was 340 Hz, and the measured noise at this condition was 70 dB(A).

Figure 6. Performance of the Unger and Novotny [24] compressor.

Figure 8. The model 50 compressor inside the test section of the hot-gas cycle calorimeter. Table 1. General compressor dimensions. Model 25 50

Figure 7. The linear compressor of Wang and Tai [25]. Bradshaw et al. [26] developed a comprehensive mathematical model for simulating a linear compressor for electronics cooling. The model included particular sub-models for the flow through the valves, flow leakage, motor losses, cylinder heat transfer, and piston dynamics. A compressor

Diameter (mm) 21 50

Length (mm) 100 160

Weight (kg) 0.2 1.2

Over the past decade, the improvement of microfabrication techniques has enabled the design and construction of new types of micro-compressors [30-37]. Shannon and coworkers [30, 31] developed a micro-miniature cooler from polymer composites via thin film layered-manufacturing techniques. Each active cooler consisted of a square patch of approximately 100 mm side and 2.5 mm thickness, and has been designed to produce a 3-W cooling capacity while operating with R-134a in a standard Rankine refrigeration cycle between 20°C (evaporating temperature) and 50°C

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(condensing temperature). The coolers could be interconnected to form a flexible refrigeration circuit. The compressor in each cooler was a touch-diaphragm double-cavity capacitive micromachine (see Fig. 10) with dimensions of 20 mm x 20 mm x 2 mm and an operating frequency of up to 300 Hz.

(a)

cycle, but were more significant in relation to the compressor, where fabrication techniques were, at the time of the study, still under development. The micro-miniature cooler technology has recently re-emerged with some novel developments which can make it also potentially applicable in the field of electronics cooling [38]. Chow and co-workers [32,33] presented the preliminary design of a mesoscale vapor compression cooling system for two main applications: (i) cooling of electronics and photonic chip modules and (ii) personal cooling. The heart of the system, shown in Fig. 11, is a high-speed R-134a centrifugal compressor with overall dimensions and characteristics that vary according to the application. For electronics cooling, the impeller outside diameter was designed with a 1.2 cm and a frequency of 400,000 rpm. Mechanical and electrical design aspects of the systems were discussed extensively, but no prototype has been built, according to the present author’s best knowledge. The type of motor to drive the compressor can differ according to the application; variable capacitance (synchronous) and electromagnetic motors have been evaluated.

(b)

Figure 9. Performance parameters of the Model 50 linear compressor as a function of evaporating and condensing temperatures. Figure 11. The mesoscale refrigerator of Chow and co-workers [32,33].

Figure 10. Cross-section view of the touch-diaphragm compressor [30]. Shannon [30] discussed both technical and manufacturing obstacles associated with the integrated miniature cooler circuit technology. These were related to all components of the

Hao et al. [34] presented the design and discussed the basic working principle of a miniature reciprocating compressor driven by an electrostatic comb drive. The discshaped compressor was 6 mm in and 0.5-mm thick. The compression chamber consists of a central boss surrounded by a ring-shaped corrugation diaphragm (driven by the combdrive) and two active micro-valves. With a working voltage of 320V, the device gave air flow rates of 2.7 ml/ min and a pressure ratio of 1.7. More recently, Sathe et al. [35-37] presented a detailed analysis of an electrostatically actuated microcompressor for electronics cooling. The compressor consists of a flexible circular diaphragm clamped at its circumference (Fig. 12.a). When a DC voltage is applied between the diaphragm and one of the chamber halves, the diaphragm is displaced as a result

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of an electrostatic attraction force. As the force is inversely proportional to the distance between the chamber surface and the diaphragm, a progressive zipping of the diaphragm starts from the edge toward the center of the chamber. The working principle of the electrostatic diaphragm compressor was evaluated against experimental data and independent models from other authors for electrostatic-actuated pumping devices.

(a)

(b)

Figure 12. The miniature diaphragm compressor of Sathe et al. [35-37]. (a) Compressor geometry. (b) Forces acting on the diaphragm.

THERMODYNAMIC ANALYSIS AND ASSESSMENT OF THE TECHNOLOGIES

CRITICAL

Refrigerant and Cycle Design R-134a has been, by far and large, the most widely used refrigerant in small-scale refrigeration applications (e.g., electronics and personal cooling). This is due to its well-know physical properties, safe handling (non-flammable, non-toxic), low cost and compatibility with many products and parts already available in the market. For cooling of electronics, since R-134a is not dielectric, some form of electrical insulation is needed between the refrigerant and the circuits or microprocessors. Except for the studies of Coggins and co-workers [20, 21] and Wadell [22], only single-stage cycles have been

investigated for electronics cooling. The main advantage of a two-stage cascade system is that lower evaporating temperatures can be achieved at moderate (i.e., around atmospheric) evaporating pressures by using a low normal boiling temperature refrigerant in the low-stage cycle. However, this benefit comes at the expense of operating with an intrinsically less efficient, more expensive and potentially noisier cycle due to the operation with two compressors and an inter-stage heat exchanger. Phelan et al. [39] performed a comparison of four alternative system designs and concluded that the cycle with the smallest energy consumption for their application was a cascade system in which the electronic device was cooled by a pumped loop that is connected to a conventional vapor compression cycle. This concept is also being pursued in personal cooling applications, whereby the refrigerant cools a secondary fluid (water) that comes into contact with the human body in a cooled garment [40]. More recently, Marcinichen et al. [41] proposed three alternative cycles (one with a pump, one with a compressor and a hybrid of the two) for cooling computer blades in highend servers. The hybrid cycle is characterized by the interchangeability between the first two cycles. The authors suggested that the best cycle for a specific application should be decided based on the overall cycle efficiency, which is defined as the ratio of the recovered energy in the condenser and subcooler to the energy required to pump the refrigerant. This enables a more integrated design, which will lead to a more economic operation of high performance data centers. Ribeiro et al. [42] introduced a hybrid cooling system in which the evaporator of a miniature refrigerator with characteristics similar to those of Mongia et al. [9] is connected to the condenser region of the heat pipes used for chip cooling. A model for the evaporator/heat pipe assembly was proposed, and an evaporator prototype was designed, fabricated and tested with R-600a at saturation temperatures of 45 and 55°C, mass flow rates between 0.5 and 1.5 kg h -1 and heat transfer rates between 30 and 60 W. It can be argued on the grounds of thermal performance alone that the evaporator/heat pipe assembly is feasible only if the overall thermal resistance from the chip surface to the evaporator fluid is less with the heat pipe in place. However, as will be seen later, there are several issues related with the implementation of the vapor compression technology in portable computers which are driven by aspects such as compressor and fan noise, vibration, robustness and reliability, refrigerant leakage etc. While these issues are not fully addressed and the associated technical challenges not duly overcome, the proposed heat rejection assembly may act as a possible solution before a more definite one becomes available. Simpler types of expansion devices, like capillary tubes and orifices, have been preferred by the majority of the studies reviewed here. The large cooling capacity systems of Schmidt and Notohardjano [15] and Agwu Nnanna [18] have employed thermostatic expansion valves, which permit some control of

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the refrigerant flow rate under varying heat loads and during start-up. Kulkarni et al. [2] discussed the effect of thermostatic expansion valve parameters (i.e., size, thermal capacities and location) on the stability of a refrigeration system. Quantitative Analysis Figure 13 illustrates the relationship between the system volume and its cooling capacity [1]. Some references mentioned in the legend of the figure are reviewed in detail in Barbosa et al. [1]. They were not described here for not being directly related to electronics cooling. Apart from the works of Bash et al. [17], Wadell [22], Coggins [20)] and Mongia et al. [9], there seems to be a direct relationship between the two variables, with the cooling capacity increasing with the system volume.

used instead. This was the case for the works of Maveety et al. [16], Mongia et al. [9] and Coggins [20]. A decreasing trend can be observed for the volumetric cooling capacity as a function of the temperature difference. The main deviations from the trend line are due to Maveety et al. [16] and Mongia et al. [9]. While the former deviation is due to an underestimation of the real temperature difference, the latter can be attributed to an underestimation of the system volume. Therefore, in broad terms, Figs. 13 and 14 suggest that a general scaling rule with respect to the system volume has been followed by the miniaturized refrigeration systems proposed in the literature.

Figure 14. Cooling capacity per unit volume as a function of the difference between the condensation and evaporation temperatures [1]. Figure 13. Relationship between the volume of the refrigeration system and the cooling capacity [1]. For the works of Wadell [22] and Coggins [20], the departure from the suggest data trend line can be explained by larger system volume associated with the two compression stages. For the work of Mongia et al. [9], the departure is possibly due to the fact that the volume reported by the authors is the internal fluid volume instead of the overall system volume, as for all the other works. From the picture of the experimental facility of Mongia et al. (Fig. 4), it is likely that the volume of the components of the system is of the order of 102 cm3 (after all, they should be able to fit into a notebook computer), which would bring the data point closer to the trend line. In Fig. 14, the cooling capacity per unit volume is plotted against the temperature difference between the condenser and the evaporator (temperature lift) reported by each study [1]. For systems where the Tcond data were not available, Tamb was

Barbosa et al. [1] carried out a comparison between the performances of miniature vapor compression systems on the basis of the refrigeration (or second-law) efficiency [43], which is defined as the ratio of the system COP and the COP of a Carnot refrigerator operating between Tamb and Tj. Thus:

R 

COPR COPC

(4)

where: COPR 

COPC 

9

q w

(5) Tj

Tamb  T j

(6)


The refrigeration efficiency can be broken-up into two efficiencies that account for the internal and external irreversibilities as follows:

 R   I E 

COPR COPI COPI COPC

(7)

where COPI is the coefficient of performance of a Carnot refrigerator based on the saturation temperatures of the refrigerant in the condenser and evaporator. Thus: COPI 

Tevap

(8)

Tcond  Tevap

External irreversibilities are associated with the transfer of heat across finite temperature differences in the evaporator and condenser and with the superheating of the compressor discharge vapor. In electronics cooling, air is the natural heat rejection medium at the condenser. To minimize the thermodynamic losses, the incoming air temperature should be as close as possible to the ambient temperature. This may be an issue if the condenser is confined to the inside of a computer or a mounting rack, where the temperature is likely to be higher than that of the external environment. Internal irreversibilities take place as result of: (i) Entropy generation in the irreversible throttling process; (ii) Pressure drop due to friction in the evaporator, condenser and connecting lines. This gives rise to more compressor work as the vapor reaches the compressor suction line with a larger specific volume; (iii) Compressor losses. Compression is no longer isentropic, and electrical, friction and thermodynamic losses are combined to reduce the compressor isentropic efficiency; (iv) Refrigerant superheating. Although it generally represents a decrease in COP (due to an increase in

compressor work), in some systems, refrigerant superheating at the evaporator outlet is desirable in order to prevent the carryover of liquid droplets into the compressor. Combined liquid subcooling and vapor superheating can be achieved in an internal heat exchanger (also referred to as suction line heat exchanger). The most common configuration of this heat exchanger is achieved by welding the capillary tube onto the wall of the compressor suction line [44]. To the best knowledge of the present author, the internal heat exchanger has not been adopted in any of the small-scale cooling systems proposed so far. The data provided by each reference gathered in this and other reviews [1] are not sufficiently complete to enable the calculation of external irreversibilities associated with the evaporator/heat sink. Only the works of Trutassanawin et al. [12], Ernst and Garimella [40] and Wu et al. [45] — the latter two being systems proposed for personal cooling — permit the calculation of the external irreversibilities in both condenser and evaporator. Information regarding the temperature of the cold source (junction, heat sink, copper block or chilled water, depending on the particular reference) is lacking in all but these three papers. Therefore, COPC could not be evaluated in some references. Table 2 [1] summarizes the calculated internal, external and refrigeration efficiencies of the published works in which at least one of such efficiencies could be computed. The values of COPI reported in Table 2 correspond the evaporating and condensing temperatures also reported in the table. The values of COPR have been obtained from each reference for each particular operating condition. In the works of Bash et al. [17], Mongia et al. [9] and Agwu Nnanna [18], the refrigeration efficiency is very likely to decrease with respect to ηI due to the external irreversibilities. In the conditions explored by Trutassanawin et al. [12], even under active cooling with the refrigeration system, the temperature of the heater is higher than that of the ambient. Although it has been demonstrated that their system is capable of keeping the heated surface temperature below the maximum allowed temperature of 85oC, it does not make sense to calculate COPC

Table 2. Summary of efficiency parameters of selected refrigeration systems [1].

Mongia et al. [9] Truttasanawin et al. [12] Bash et al. [17] Agwu Nnanna [18] Ernst and Garimella [40] Wu et al. [45]

Tevap [oC] 50 21 25 4 24.9 7

Tcond [oC] 90 60 [*2] 52 68 51.5 40

Tamb [oC] 50 27 22 20 43.5 40

Tj [oC] n.d.p.[*1] 78 n.d.p. n.d.p. 29.9 18.73

COPR [-] 2.25 4.0 4.0 1.2 5.0 2.3

COPC [-] 22.28 13.72

COPI [-] 8.08 7.54 11.04 4.33 11.20 8.49

ηI [-] 0.28 0.53 0.36 0.28 0.45 0.27

ηE [-] 0.50 0.62

ηR [-] 0.22 0.17

[*1] n.d.p.: no details provided, [*2] inferred from their nominal operating data.

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for their conditions. The values of the internal efficiency reported by Trutassanawin et al. [12] ranged from 33% to 52%. FINAL REMARKS AND RECOMMENDATIONS This paper presented an assessment of vapor compression refrigeration for cooling of electronics. A significant number of system prototypes and compressors have been presented in the literature. The main conclusions arising from this study are as follows: 1. Mechanical vapor compression refrigeration seems to be a viable technology for application in personal and electronics cooling. A variety of systems with different features and intended for diverse applications (e.g., computer servers, desktop, notebook computers etc.) have been proposed. However, there are a number of technical challenges that need to be overcome in order to make the technology widely available in the market for these and other applications. Some technical challenges are as follows: a. Compressor reliability and high efficiency when at the small scale. This issue is closely linked to the thermal management of the compressor, which is crucial for guaranteeing its long life. In a compressor, a significant fraction of the energy input is converted into heat. With the decreasing size of the compressor unit, there is progressively less surface area available for heat dissipation, which causes a rise in temperature at several points of the compressor. The temperature rise lowers the compressor efficiency, thereby increasing its power consumption. Integrated design solutions need to be found in order to keep the overall system size as small as possible; b. Reduced noise and vibration. Although compressor noise may not be a relevant issue in some types of personal cooling applications, it can certainly be of great concern for some electronics cooling devices, especially because the majority of the miniature compressors have been design to operate at fairly high frequencies. System reliability issues associated with vibration-induced fatigue caused by the compressor is also an important aspect to be considered, since it may result in refrigerant leakage. c. Ability to handle varying workloads, power dissipation surges and fast transients. The dynamic behavior of small-scale refrigeration systems under transient conditions (e.g., compressor start-up and shut down) should be investigated in a systematic manner [18].

2.

3.

4.

Effective control strategies [46,47] have to be developed to deal with the complex interaction between the compressor and the expansion device under variable thermal loads. Moreover, the refrigeration system needs to be able to minimize spatial temperature gradients on the low-temperature source (heat sink). d. Availability at a low cost. This is related primarily to the use of inexpensive materials and refrigerants. However, cost may not be a too serious issue for some high-end and specific applications. e. Moisture condensation management in the heat sink region (for electronics cooling). The challenge here is to take advantage of the below-ambient temperature potential of mechanical vapor compression whilst maintaining the electrical integrity of the heat source. A sort of scaling rule has been observed when the system volume was plotted against the cooling capacity. The cooling capacity per unit system volume also followed a clear decreasing trend with respect to the system temperature lift. The trends seemed not to be influenced by the compressor type. Several small-scale compression technologies have been investigated in the open literature. In the past four years or so, special attention has been devoted to the development of rotary and linear compressors for miniaturized refrigeration applications. These technologies seem, at the present time, more mature than the smaller compressors (e.g., electrostatically actuated) also reviewed in this paper. From an academic standpoint, more detailed experimental data are needed to allow a comprehensive evaluation of the sources of thermodynamic losses in each component of the system, more notably, the heat sink/evaporator and the compressor.

ACKNOWLEDGEMENTS The author is indebted to CNPq through Grant No. 573581/2008-8 (National Institute of Science and Technology in Cooling and Thermophysics). Support from P. A. de Oliveira (UFSC), G. B. Ribeiro (Embraco), L. W. da Silva (Embraco), F. H. Klein (Embraco), P. R. C. Couto (Embraco), A. Santos (UFSC) and C. Melo (UFSC) is greatly appreciated. REFERENCES [1] Barbosa Jr., J.R., Ribeiro, G.B. and Oliveira, P.A., A Stateof-the-Art Review of Compact Vapour Compression Refrigeration Systems and their Applications, Heat Transfer Engineering, in press, 2011. [2] Kulkarni, A., Mulay, V., Agonafer, D. and Schmidt, T., Effect of the Thermostatic Expansion Valve

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