Abstract
Microchannel heat sink (MCHS) is an advanced cooling technique to fulfil the cooling demand for electronic devices installed with high-power integrated circuit packages (microchips). Various microchannel designs have been innovated to improve the heat transfer performance in an MCHS. Specifically, the utilisation of nanotechnology in the form of nanofluid in an MCHS attracted the attention of researchers because of considerable enhancement of thermal conductivity in nanofluid even at a low nanoparticle concentration. However, a high-pressure drop was the main limitation as it controls the MCHS performance resulted from heat transfer augmentation. Therefore, this study aimed to critically summarise the challenges and limitations of both single and hybrid passive methods of MCHS. Furthermore, the performance of nanofluid as a coolant in the MCHS as affected by the type and concentration of nanoparticle and the type of base fluid was reviewed systematically. The review indicated that the hybrid MCHS provides a better cooling performance than MCHS with the single passive method as the former results in a higher heat transfer rate with minimal pressure drop penalty. Besides that, further heat transfer performance can be enhanced by dispersing aluminium dioxide (Al2O3) nanoparticles with a concentration of less than 2.0% (v/v) in the water-based coolant.
1 Introduction
Application of cooling system in thermal engineering is recognised and have been studied both theoretically and practically in building energy system, electronic devices, chemical vapour deposition instruments, solar energy collectors, furnace engineering, and many more [1]. In recent years, thermal management of electronic devices is of interest as a new generation of high performing dense chip packages that function at high frequency produces a very high heat flux on the electronic devices. Prolonged heat flux creates a hot spot on the electronic devices and reduces the lifespan of the electronic devices because of the acceleration of the mean time to failure (MTTF) as described by the Black’s equation [2]. Thermal management of compact electronic devices that operate at high power density is critical as there is a lack of efficient technique to remove heat dissipation from the electronic devices [3]. The increase in power density and miniaturisation of electronic packages have changed the direction of cooling system technology, from air-cooling technology to advanced heat transfer technology because of the inadequacy of conventional method to remove extreme high heat flux [4].
Several methods have been proposed in the previous studies to improve the cooling performance, such as air cooling method [5], heat pipe method [6], use of liquid material as a coolant [7], and micro-cooling method [8]. Micro-cooling is a good technique because of its high cooling efficiency compared to the other methods. Even though the air cooling method is the simplest cooling technique, it has a low cooling efficiency with additional heat generation by the fan itself [9]. Meanwhile, for the liquid material and heat pipe applications, space is required to accommodate additional system for condensation process [9], which is not suitable for compact electronic devices. Therefore, the micro-cooling method is one of the promising techniques that can dissipate a very high heat flux generated by the compact electronic devices that can be attributed to the high heat transfer surface-area-to-volume ratio. Besides that, the micro hydraulic diameter also contributes to the heat transfer performance enhancement because of the augmentation of heat transfer coefficient [10]. When the micro-cooling method was used, microchannel heat sink (MCHS) was found to be the most prospective as a very high heat flux can be removed compared to micro-jet impingement, micro-heat pipe, and micro-electro-hydrodynamic methods [11]. In 1981, Tuckerman and Pease [12] reported that rectangular MCHS can remove heat-flux up to 790 W/cm2. However, the pumping power required by the MCHS was very high because of high-pressure drop penalty generated in the microchannel. Based on the study of Japar et al. [13], high-pressure drop penalty was attributed to high wall shear stress in the developing region of laminar flow. The discovery motivated many scholars to investigate the performance of rectangular microchannel via geometry parameter and stacked layer optimisations.
Khan et al. [14] stated that when the aspect ratio of a rectangular microchannel (channel width to channel height) increased, thermal and hydraulic resistances decreased because of the larger convective heat transfer area and flow cross-section area, respectively. The findings were further supported by Kowsari et al. [15] where for a smaller and constant cross-section area, the heat transfer performance increased with channel aspect ratio. In-depth effect of a rectangular channel aspect ratio on the thermal performance was investigated by Sobhan et al. [16] and Kou et al. [17]. The authors found that at a fixed aspect ratio, channel width significantly affected thermal resistance as the bottom channel area lied directly under the applied heat flux. Besides the thermal resistance issue, the pressure drop penalty was also an aim of rectangular channel MCHS optimisation process. One of the related studies includes investigation of hydrothermal performance in a single-layered and double-layered MCHS to reduce the pressure drop penalty by Chong et al. [18]. The study reported that in the laminar region, the double-layer MCHS reduced the pressure drop by 53.9% compared to the single-layered MCHS. However, the thermal resistance in the double-layered MCHS was higher than the single-layered MCHS by 20.8%. A solution was proposed by Wei and Joshi [19] to reduce the thermal resistance in a stacked MCHS, either by increasing the pumping power or by reducing the channel length. Hung and Yan [20] also recommended increasing the pumping power to mitigate the optimal thermal resistance. However, in another novel study by Wang et al. [21], both thermal resistance and pumping power consumption were improved simultaneously through optimisation of semi-porous rib geometry. The researchers reported that the double-layered MCHS with the optimised semi-porous-rib geometry improved the cooling performance and pumping power by 14.06% and 16.40%, respectively.
Nevertheless, further geometry optimisation of a rectangular channel MCHS was limited by pressure drop penalty in a monolayer MCHS and additional space requirement by a multilayer MCHS in the cooling system of a compact electronic device. In recent years, nanotechnology was applied to further improve the performance of hydrothermal. According to Nanotechnology journal database, nanotechnology is widely used in various fields such as medical [22,23,24], biomedical especially for cancer detection [25,26,27] and treatment [28,29], bone regeneration [30,31,32], vaccine [33] and drug [34] delivery system, central nervous system [35], biological imaging [36], and membrane proteins analysis [37,38]. However, studies on the nanofluid application for thermal management [39] are limited.
Recently, many investigations were conducted to enhance the performance of MCHS through advanced geometric structure and nanofluid [40], which can provide high thermal performance with an acceptable pressure drop. Adham et al. [41] presented a review on the effect of constant cross-section shape on hydrothermal performance to provide a better understanding of research related to geometric structure. Simple cross-section channels such as rectangular, trapezoidal, triangular, and square shapes were reviewed to find a suitable flow condition and performance indicator for MCHS application. In another review by Kumar et al. [42], nanofluid behaviour in various shapes of MCHS was discussed. Based on both the reviews, the best shape of MCHS that has a higher thermal performance and a lower friction loss was suggested. However, the difference in performance between a hybrid MCHS (the combination of more than one passive method) and single passive method of MCHS was not included in both the reviews. Besides, the range of nanoparticle concentration, type of nanoparticle, and base fluid that are commonly used in MCHS application were also not discussed in both reviews, and hence the details are presented in the current review. Therefore, the challenges and limitations of single passive and hybrid passive methods that were adopted in microchannel designs are discussed critically in the present review based on the respective geometry structure and mechanism.
2 Hydrothermal performance enhancement technique
For the past 30 years, many methods were proposed to improve the overall performance of MCHS with minimal thermal resistance and pressure drop that can satisfy the cooling demand. Generally, the methods can be categorised into two groups namely active and passive. The active method uses external energy in the system to improve the heat transfer performance, whereas the passive method does not require external energy. In general, the passive method was preferred because of its low cost and absence of mechanical moving part compared to the active method. Figure 1 shows the techniques used as part of the active method [43,44,45,46,47,48,49] and passive method [50,51,52,53,54,55,56,57,58,59,60,61,62,63,64,65,66,67,68,69,70,71,72,73,74,75,76,77,78,79,80,81,82,83,84,85,86,87,88] that are available in the open literature.
Enhancement technique for MCHS.
2.1 Active method
Since the middle of 1981, majority of studies emphasised on the passive method than the active method in the performance enhancement of MCHS because of the compact design of advanced electronic devices. However, few researchers attempted to improve heat transfer performance using the active method. In 2003, Go [43] evaluated the effect of flow-induced vibration of a microfin array on heat transfer augmentation. The study reported that the high heat transfer rate was obtained by increasing the vibration displacement of the microfin. In another innovative work, Krishnaveni et al. [46] proposed a periodic electric field technique in rectangular MCHS where chaotic mixing in a microchannel was induced which contributed to heat transfer augmentation. In a different study, Hessami et al. [47] reported on the effect of flow pulsation technique on microchannel performance and concluded that heat transfer enhancement was obtained by increasing the frequency and reducing the amplitude of flow pulsation. Variable roughness structure technique is not popular and least studied, which could be because of the complexity of the system designed for the compact area.
2.2 Passive method
2.2.1 MCHS with single/individual passive method
Single/individual passive method of MCHS refers to a design that implements only one of the passive methods in its channel design with less complex geometry structure in contrast to the hybrid design MCHS. However, this single passive method of MCHS is only efficient at low Reynolds number (Re) because of high-pressure drop penalty at higher Re as shown in Table 1. There are three main factors that contribute to the enhancement of the heat transfer performance in the single passive method of MCHS, namely high degree of fluid mixing, mainstream flow separation, and the periodic break-up of the boundary layer (Table 1). These flow mechanisms can be caused by the Dean vortices formation, secondary flow generation, jet and throttling effect, and the entrance effect. It is worth stating here that the aforementioned flow behaviours assist in the enhancement of heat transfer performance besides the convective heat transfer area factor.
Performance of MCHS with single/individual passive method
| Flow disruption | ||||
| Author: Hong and Cheng (2009) [57] | Geometry structure: Offset strip-fin microchannels | Re: 75 | ||
| Optimum geometry parameter: The ratio of fin interval to fin length is 1 | ||||
| Observations: (1) The heat transfer coefficient increases with the number of fins as the flow disturbance and break-up of the boundary become more frequent; (2) heat transfer area increases with the number of fins; (3) lower flow rate is required as a higher number of fins provides a higher heat transfer rate; (4) the optimum number of fins is vital to prevent the higher pressure drop penalty | ||||
| Author: Chai et al. (2013) [58] | Geometry structure: Interrupted microchannel with rectangular ribs in the transverse microchambers | Re: 187–715 | ||
| Optimum geometry parameters: For Re < 400: Rectangular rib width (w): 0.14 mm; rectangular rib length (l): 0.5 mm; the distance from the parallel microchannels to the rib row, d = 0.2 mm and 0.25 mm; the space between two adjoining transverse microchambers, s = 1.3 mm | ||||
| Observations: (1) The rib length and width have a significant effect on heat transfer performance; (2) the friction factor and heat transfer rate increased with the length and width of ribs geometry; (3) the friction factor augmentation was more affected by the rib width than the rib length; (4) the friction factor and heat transfer rate were more affected by the space between two adjacent transverse microchambers, s, than the distance from the parallel microchannels to the rib row, d | ||||
| Author: Chai and Wang (2018) [61] | Geometry structure: Interrupted MCHSs with different rib geometries in transverse microchambers (rectangular, backward triangular, diamond, forward triangular, and ellipsoidal ribs) | Re: 187–715 | ||
| Optimum geometry parameter: The ellipsoidal ribs with the length of 0.5 mm showed the best overall performance | ||||
| Observations: (1) Ribs geometry with a flat surface at the leading edge perpendicular to the mainstream produced a high friction factor; (2) the effect of rib length on the friction factor augmentation depends on the shape of ribs geometry | ||||
| Author: Xie et al. (2014) [59] | Geometry structure: MCHSs with different multistage bifurcations | Re: 230–560 | ||
| Optimum geometry parameter: Single-stage bifurcation with long bifurcation plate (length: 25 mm; width/thickness: 35 µm) | ||||
| Observations: (1) Uniform temperature distribution at the heated wall; (2) pressure drop augmentation in the MCHS with multistage bifurcations was affected by bifurcation plates length close to the channel exit; (3) two-stage bifurcations were the optimum stage number in the MCHS with multistage bifurcation design; (4) the pressure drop augmentation increased with the number of bifurcation plate because of the enhancement of fluid velocity in the smaller channel | ||||
| Channel curvature | ||||
| Author: Sui et al. (2010) [66] | Geometry structure: Wavy microchannel with the rectangular cross-section | Re: 100–800 | ||
| Optimum geometry parameter: Wavy microchannel with a relative wavy amplitude from 0.05 to 0.125 | ||||
| Observations: (1) The heat transfer performance and pressure drop penalty increased with increasing relative wavy amplitude of microchannel; (2) the optimum relative amplitude of the wavy channel is essential to prevent high-pressure drop penalty that dominates the heat transfer performance augmentation; (3) at a fixed relative wavy amplitude, the number of Dean vortices formation increased with the Re; (4) stretching and folding of fluid-particle generated by several Dean vortices promoted a higher degree of flow mixing | ||||
| Author: Mills et al. (2016) [67] | Geometry structure: Microchannel with the wavy-walled channel | Re: Re < 800 | ||
| Optimum geometry parameters: For 280 < Re < 385: (1) The ratio of wall amplitude to the distance between sidewalls, A, was 0.35; (2) the ratio of wall oscillation period to the distance between the sidewall and the centre line of the microchannel, L, was 3. For Re > 385: (1) The ratio of wall amplitude to the distance between sidewalls, A, was 0.25; (2) the ratio of wall oscillation period to the distance between the sidewall and the centre line of the microchannel, L, was 3 | ||||
| Observations: (1) The amplitude of wall waviness was more effective at a lower flow rate in a steady regime for heat transfer performance enhancement; (2) heat transfer performance augmentation in an unsteady regime was higher than in a steady regime; (3) heat transfer enhancement in an unsteady region was influenced more by the pressure drop than the wall amplitude; (4) the formation of Dean vortices in an unsteady regime was more effective than in a steady regime for heat transfer performance enhancement | ||||
| Author: Lin et al. (2017) [68] | Geometry structure: MCHS with a wavy channel | Re: 300–800 | ||
| Optimum geometry parameters: Two adjacent wavy geometries with the wavelength and amplitude difference of 400 and 8 µm, respectively | ||||
| Observations: (1) The amplitude and wavelength of wavy geometry have a significant impact on the thermal resistance and bottom wall temperature; (2) low thermal resistance and small temperature difference at the bottom wall can be obtained by increasing the wave amplitude and reducing the wavelength of wavy geometry; (3) greater Dean vortices formation at the channel inlet can be obtained by increasing the wavy wavelength and reducing the wave amplitude; (4) greater Dean vortices formation at the channel outlet can be obtained by reducing the wavy wavelength and increasing the wave amplitude; (5) superior thermal performance can be achieved by increasing the wavelength or amplitude difference between two adjacent wavy geometries with small aspect ratio | ||||
| Secondary flow | ||||
| Author: Lee et al. (2012) [74] | Geometry structure: Microchannel with oblique fins | Re: 300–800 | ||
| Optimum geometry parameter: The oblique angle was 27° | ||||
| Observations: (1) Diversion of the mainstream to a small fraction of flow into adjacent main channels; (2) a thinner boundary layer zone formation was related to fluid fraction flow from the main channel to the secondary channel; (3) the oblique fin geometries increased the heat transfer area and subsequently enhanced the heat transfer rate which also increased the pressure drop; (4) the thermal boundary layer decreased with increasing Re; (5) the existence of a secondary channel provided by the oblique geometry further improved the heat transfer performance by disrupting the mainstream of adjacent main channel; (6) high pumping power was required to divert a greater amount of flow into the secondary channel with a stronger momentum which increased the heat transfer performance with excessive pressure drop penalty noted; (7) smaller fin pitch increased the intensity of boundary layer development; (8) smaller oblique angle increased secondary flow into the secondary channel | ||||
| Author: Lee et al. (2015) [77] | Geometry structure: Microchannel with oblique fins | Re: 180–680 | ||
| Optimum geometry parameters: Hydraulic diameter, oblique angle, and fin pitch were 104 µm 27° and 400 µm, respectively | ||||
| Observations: (1) The thermal resistance in the enhanced microchannel decreased rapidly with the increase in Re; (2) oblique fin geometries provided a more uniform temperature distribution but increased the pressure drop; (3) the high-pressure drop was generated after Re achieved 500 | ||||
| Author: Kuppusamy et al. (2014) [75] | Geometry structure: Microchannel with the slanted secondary channel in alternating orientation | Re: N/A | ||
| Optimum geometry parameters: The secondary channel width (α), length of the longer edge of the fins (β), and secondary channel angle (θ) were 64, 160 µm, and 37°, respectively | ||||
| Observations: (1) Presence of secondary channels in alternating orientation enhanced the heat transfer performance, besides reducing pressure drop in the enhanced microchannel; (2) heat transfer rate increased with the width of the secondary channel; (3) thermal resistance and pressure drop decreased with increasing secondary channel width; (4) heat transfer rate increased with decreasing longer edge length of fins and secondary channel angle; (5) thermal resistance and pressure drop decreased with a longer edge length of fins and secondary channel angle | ||||
According to Bhandari and Prajapati [89], pin fin height affects the heat transfer performance because of Dean vortices formation. The study reported that the enhancement of heat transfer performance in the microchannel with a fin height of 1.5 mm (convective surface area: 639 mm2) was higher when compared with a fin height of 2.0 mm (convective surface area: 687 mm2). The vortices formation increased the effective area of convective heat transfer. In another novel study [90], the formation of Dean vortices in microchannels was reported to reduce the boundary layer thickness and provide a more uniform velocity distribution along the microchannels.
Among the passive methods, surface roughness [50,51,55] and re-entrant obstructions [72] got less attention by previous researchers. Geometry parameter in the surface roughness method is difficult to control during fabrication process to get the uniform surface roughness. The relative height of roughness affects the size of the vortex flow formation adjacent to the roughness element. The larger the vortex flow formation size, the weaker the heat transfer performance [55]. Besides that, the surface roughness method is only more effective at the higher heat flux [51], fluid temperature, and heating power [50]. For the re-entrant obstructions method, high-pressure drop penalty is produced because of the entrance effect influenced by constriction–expansion geometry [72].
2.2.2 MCHS with hybrid passive method
Numerous studies and reviews were carried out to explore the potential of a single passive method for microchannel design innovation compared to hybrid passive method because of the less complex geometry of the former. However, high-pressure drop issue is the main constrain of single passive method of MCHS. To minimise pressure drop and increase the heat transfer performance, hybrid passive method of MCHS which use more than one passive method in its channel design is of growing interest. Recently, many researchers have used the hybrid passive method in a single-phase flow for the enhancement of MCHS performance. In 2020, van Erp et al. [91] reported a sustainable enhanced MCHS with three important solutions: (1) high cooling performance; (2) less water and energy consumption; and (3) small and compact heat sink. Li et al. [92] proposed a novel high-efficiency-low-resistance heat exchanger design. The study reported that the combination of the dimple and pin-fin geometries used in the design provided a higher heat transfer rate with energy-saving and low resistance features enhanced the thermal performance by 10.3% at a Re of 200. Besides, the area of the high-temperature region and temperature gradient decreased. Furthermore, the violent action of flow on the heated wall provided by pin-fin increased the uniformity of the heated area.
In 2013, Gong and Wei [93] analysed the performance of microchannel structured by dimple and wavy shapes where the effect of dimple number in the wavy wall was studied numerically. The presence of the dimple structure in the wavy channel was reported to enhance the heat transfer performance but not the flow resistance. Besides, the overall performance of microchannel was enhanced with increasing dimple number. In a subsequent study in 2016, Gong et al. [94] investigated similar geometry combination with different orientation of dimple structure. The experimental findings showed that microchannel with dimple structures in the throat of the wavy channel has the best thermal performance compared to dimple structure in the cavity of the wavy channel. The dimple structure in the cavity area reduced the overall performance provided by the wavy channel as the velocity gradient reduced on the contact surface.
A year later, in 2017, the performance of convergent-divergent channel with ribs and cavities on the sidewall was reported by Srivastava et al. [95] for constant heat flux and Re of 120–900. The findings showed that the developed design reduced the overall thermal resistance up to 40% and the temperature of the microchannel bottom surface was relatively uniform. The combined effect of interruption, redevelopment of the boundary layer, and recirculation zone increased the average Nusselt number from 15% to 46% which contributed to the heat transfer augmentation. However, the overall performance of MCHS decreased when the flow rate was increased because of high-pressure drop penalty.
In recent years, the combined effects of cavities and ribs, on the sidewalls and the centre of the straight channel respectively, have gained more attention. Li et al. [96] presented a numerical study to investigate the effects of ribs and cavities on fluid flow and heat transfer characteristics for the Re of 173–635. Thermal enhancement factor and entropy generation number augmentation were used as the performance indicator. The analysis revealed that combined effect of interruption, redevelopment of the thermal boundary layer, intensified mainstream disturbance, and chaotic mixing between hot and cold water enhanced the heat transfer performance. After a year, Li et al. [97] extended their research using various shapes of ribs and cavities in a straight microchannel, and the rib shapes were found to influence thermal and hydraulic performances to a greater extent. Meanwhile, the triangular ribs showed the lowest heat transfer performance compared to other designs, although it has the lowest pressure drop penalty because of delayed separation point of the mainstream on the triangular geometry. In contrast, the rectangular rib design has the highest pressure drop penalty among all the designs compared. In the same year, Ghani et al. [98] reported that heat transfer augmentation with a minimal pressure drop in a microchannel with sinusoidal cavities and rectangular ribs can be achieved by geometrical optimisation. The optimum relative cavity amplitude, rib width, and rib length found in the study were 0.15, 0.13, and 0.5, respectively. The combination of these optimum parameters resulted in the highest performance factor compared to other designs studied.
In contrast to the aforementioned design configuration, Xia et al. [99,100] analysed the effects of ribs and cavities at the sidewalls on heat transfer performance using two different methods, namely computational fluid dynamics (CFD) and CFD with multi-objective evolutionary algorithm (MOEA). In the CFD method, the analysis found that a relative rib height significantly affected entropy generation. Not only that, the combined effect of flow mixing and redevelopment of the boundary layer in the microchannel increased the heat transfer performance. From the data obtained, Xia et al. [99,100] established a correlation between Nusselt number and friction factor as shown in equations (1) to (3). Meanwhile, in CFD with MOEA method, the result illustrated that relative rib height increased, thermal resistance and pressure drop penalty decreased and increased, respectively. The relative cavity height and rib height exhibited similar performance trend except for the very deep cavities that could be influenced by the formation of a stagnant zone. Besides, MOEA results indicated a relationship between total thermal resistance and pumping power.
For a better understanding of the effect of geometry configuration or arrangement of ribs and cavities in the straight channel on the thermal performance, Xie et al. [101] developed several designs to analyse the best configuration of rib and cavity combinations. The highest Nusselt number ratio was obtained when using the configuration of ribs on both sidewalls. The Nusselt number ratio decreased according to the following combinations: cavities and ribs on opposite sidewalls > ribs on the sidewall > cavities on the sidewall. For friction factor ratio, ribs on the sidewall and ribs on both sidewalls exhibited the highest values when compared with other configurations. In general, different types of rib and cavity arrangement could enhance the heat transfer performance compared to a straight microchannel design. However, from the viewpoint of energy saving, the arrangement of the ribs and cavities is very important to prevent a high-pressure drop.
The combination of secondary channel and rib geometries is one of the novel designs which could increase the flow mixing and redevelopment of the boundary layer. The combination of the geometries was applied by Ghani et al. [102] to investigate the performance of combined effect on the fluid flow and heat transfer characteristic for the Re of 100–500. The key strategies that were considered in developing a novel design include reduction in pressure drop, the increment of flow mixing, and interruption of main flow for redevelopment of the boundary layer. Optimum parameters for the design were suggested by the authors to achieve maximum thermal performance with a minimal pressure drop penalty.
Another novel design using impinging jet with a dimple structure showed great potential in thermal performance enhancement. Dimple structure design development of an MCHS is important as dimple could increase or decrease the overall performance of the MCHS. Huang et al. [103] studied the effect of various shapes of dimples on the fluid flow and heat transfer characteristics in an impinging jet microchannel where convex dimple showed the highest performance enhancement compared to mixed dimples and concave dimple. The presence of convex dimple structure decreased the flow resistance in the microchannel resulting in a lower pumping power consumption. In 2018, similar geometries were analysed by Ming et al. [104,105] to determine the optimum geometry parameter of the convex dimple to reduce the heat flux up to 500 W/cm2.
Overall, the studies highlight the potential of the combined technique of passive method in a single channel. It was noted that the combined effect from each technique in a passive method resulted in a significant impact on the thermal performance of an MCHS. However, the high-pressure drop issue that affected the pumping power consumption requires consideration in the design development to ensure that the overall performance of MCHS can be improved and optimised. Table 2 shows the summary of current designs [92,93,94,95,96,97,98,99,100,101,102,103,106,107,108,109,110], challenges, and limitations of MCHS with hybrid passive method.
Advanced structure, challenges, and limitations of passive methods in a microchannel
| Yan et al. [106] |
Design: Secondary flow and flow disruption methods |
|
Overall performance: The coefficient of performance ratio ((COP)r) was 0.97 at flow rate of 400 mL/min Challenges and limitations:
|
|
| Bahiraei et al. [107] |
Design: Fluid additive, secondary flow, and flow disruption methods |
Overall performance: The performance evaluation criteria (PEC) was about 1.15 at a Re of 100Challenges and limitations:
|
|
| Haziq et al. [108] |
Design: Secondary flow and flow disruption methods |
|
Overall performance: Thermal resistance decreased by 68%; pressure drop increased by 900% Challenges and limitations:
|
|
| Rui et al. [109] |
Design: Surface roughness and flow disruption methods |
|
Overall performance: Highest Nusselt number ratio and friction factor ratio of the proposed designs were about 2.8 and 9, respectively, at a Re of 600 Challenges and limitations:
|
|
| Danish et al. [110] |
Design: Flow disruption and re-entrant obstruction methods |
|
Overall performance: Thermal performance and pumping power increased by 30.6 and 11.7%, respectively Challenges and limitations:
|
|
| Li et al. [92] |
Design: Flow disruption and surface roughness methods |
Overall performance: The PEC of the proposed design was 2.12 at a Re of 300Challenges and limitations:
|
|
| Srivastava et al. [95] |
Design: Surface roughness, flow disruption, and re-entrant obstruction methods |
|
Overall performance: Overall thermal resistance decreased by 40% Challenges and limitations:
|
|
| Li et al. [97] |
Design: Flow disruption, re-entrant obstruction, and surface roughness methods |
|
Overall performance: The PEC of all analysed designs in the study were less than 1.46 at a Re of 300–400 Challenges and limitations:
|
|
| Huang et al. [103] |
Design: Flow disruption and surface roughness methods |
|
Overall performance: Highest overall performance enhancement of enhanced design was about 15% Challenges and limitations:
|
|
| Ghani et al. [98] |
Design: Flow disruption, re-entrant obstruction, and surface roughness methods |
|
Overall performance: The PEC of the proposed design was 1.85 at a Re of 800 Challenges and limitations:
|
|
| Ghani et al. [102] |
Design: Flow disruption and secondary flow methods |
|
Overall performance: The PEC of the proposed design was 1.98 at a Re of 500 Challenges and limitations:
|
|
| Xia et al. [100] |
Design: Flow disruption, re-entrant obstruction, and surface roughness methods |
|
Overall performance: The lowest thermal resistance was 0.18 K/W at the pumping power of 0.3 W Challenges and limitations:
|
|
| Li et al. [96] |
Design: Flow disruption, re-entrant obstruction, and surface roughness methods |
|
Overall performance: The PEC of the proposed design was 1.62 at a Re of 500 Challenges and limitations:
|
|
| Gong et al. [94] |
Design: Re-entrant obstruction and surface roughness methods |
|
Overall performance: Thermal resistance decreased by 60% and pumping power increased by 104% in the wavy channel with a dimple at the channel throat Challenges and limitations:
|
|
| Xie et al. [101] |
Design: Flow disruption, re-entrant obstruction, and surface roughness methods |
|
Overall performance: The range of Nusselt number and friction factor ratio were 1.45–26.19 and 0.86–110.18, respectively Challenges and limitations:
|
|
| Gong et al. [93] |
Design: Re-entrant obstruction and surface roughness methods |
|
Overall performance: The PEC of the proposed design was about 1.33 at a Re of 200 Challenges and limitations:
|
|
| Xia et al. [99] |
Design: Flow disruption, re-entrant obstruction, and surface roughness methods |
|
Overall performance: The highest Nusselt number and friction factor ratio were about 2.4 and 7, respectively, at a Re of 600 Challenges and limitations:
|
It is worth mentioning here that the combination of more than one passive methods in a microchannel design can improve or negatively affect the overall performance of MCHS. According to Lu et al. [113], the effect surface roughness of three different microchannel designs, namely wavy, dimpled, and square channels on hydrothermal performance, was highly dependent on the local flow field in each design. The surface roughness has a less impact on the dimpled and valley of the wavy channel because of low-velocity area present close to the wall.
However, in another study, Yang et al. [114] reported that the hybrid MCHS showed superior hydraulic performance over the wide range of Re (Re < 883) compared to the single passive design MCHS. Li et al. [115] also portrayed a similar trend in the overall performance of hybrid MCHS. The authors found that MCHS with isosceles triangular cavities and forward drop-shaped fins resulted in a high thermal performance improvement with a relatively small flow resistance.
Besides out of plane mixing method, secondary flow generated by a secondary channel is an alternative method, which can provide a high degree of flow mixing with a relatively small flow resistance. The effectiveness of secondary channel geometry was reported numerically in our previous research [111]. The study showed that the existence of secondary channel geometry in hybrid MCHS (TC-RR-SC MCHS) increased the degree of flow mixing between cold and hot water in the main microchannel and adjacent microchannel. Besides that, the design also resulted in the reduction in pressure drop by providing a large flow area for coolant. Model for hybrid MCHS with secondary channels is as shown in Figure 2. This design appears as the best hybrid MCHS compared to our other designs [116,117,118,119] because it increases the cooling performance, reduces the water-coolant and pumping power consumption, and small and compact in size.
![Figure 2
Hybrid MCHS with secondary channel geometry [111].](/document/doi/10.1515/ntrev-2020-0094/asset/graphic/j_ntrev-2020-0094_fig_002.jpg)
Hybrid MCHS with secondary channel geometry [111].
2.2.3 Utilisation of nanofluid in an MCHS with passive design
In the last decade, many researches were performed to innovate the design of flow passage to enhance MCHS performance. However, some studies focused on increasing the heat transfer performance by improving the transport properties such as fluid thermal conductivity in an MCHS. Continuous discovery of new coolant can improve the ability of convective heat transfer in an MCHS application. Earlier, researchers dispersed solid particle of micrometre and millimetre size in a water-based fluid to improve the thermal conductivity that can enhance the heat transfer performance. Practical limitations of the method are well established and the most common drawback was the clogging effect. In 1995, Choi and Eastman [120] introduced a new advanced potential coolant named nanofluid. Nanofluid is being used widely as a coolant additive in the MCHS application. Nanofluid is usually prepared by dispersing solid particles with nanometre size in a base fluid.
Some criteria need to be considered during the preparation of nanofluid, which include base fluid selection, size, shape, concentration, purity, dispersibility, thermal conductivity, preparation method (dispersion technique and sonication time [121,122]), and compatibility of nanoparticles that ensure the homogeneity of nanofluid [123]. Table 3 shows some types of nanoparticles that have been used in the MCHS application in the previous studies [124,125,126,127,128,129,130,131,132,133,134,135,136] such as an aluminium dioxide (Al2O3), multi-walled carbon nanotube (MWCNT), silicon dioxide (SiO2), Titanium dioxide (TiO2), silver (Ag), single-wall carbon nanotube (SWCNT), copper (Cu), copper oxide (CuO), zinc (Zn), silicon carbide (SiC), and iron(ii,iii) oxide (Fe3O4). It can be observed that most of the researchers were interested in Al2O3 nanoparticle as an additive in a base fluid for the MCHS application. Besides that, pure water such as deionised (DI) and distilled water was widely used as a base fluid in the analysis. Alfaryjat et al. [133] studied the effect of the base fluid type that was homogenised with Al2O3 nanoparticle on the fluid flow and heat transfer characteristics in a rhombus MCHS. They revealed that Al2O3 nanoparticle that was dispersed in pure water had the highest heat transfer coefficient compared to other base fluids such as engine oil, glycerine, and ethylene glycol (EG). Furthermore, Al2O3 in water provided the lowest thermal resistance among the other base fluids. Although all the base fluids have a similar friction factor, Al2O3/water generated the smallest pressure drop that minimised the pumping power consumption compared to other base fluids. The findings were further supported by Lyu et al. [137] where water-based nanofluid provided higher cooling performance and lower pumping power consumption than kerosene-based nanofluid in a fractal MCHS. For more understanding about the performance of Al2O3 nanoparticle as an additive, Thansekhar and Anbumeenakshi [128] analysed the effect of thermo-physical properties of nanoparticles on the heat transfer performance for a low Re of 55–90 experimentally. The Al2O3/water resulted in a better heat transfer performance compared to SiO2/water because of the thermal conductivity of the former. Thermal conductivity is one of the factors that determine the performance of nanofluid.
Utilisation of nanofluids in the MCHS application with different nanoparticle concentration/diameter and flow condition
| Reference | Re | Nanofluid | Volume concentration/diameter | Base fluid | Findings |
|---|---|---|---|---|---|
| Arabpour et al. [135] | 1–100 | MWCNT | 4% (v/v), 8% (v/v) | Kerosene |
Numerical work:
|
| Tran et al. [136] | 100–900 | Al2O3, TiO2 | 0.1% (v/v), 0.5% (v/v), 1.0% (v/v) | DI water |
Numerical work:
|
| Sarafraz et al. [134] | 50–1,400 | Ag | 0.01% (w/v), 0.05% (w/v), 0.10% (w/v) | DI water |
Experimental work:
|
| Alfaryjat et al. [133] | 700 | Al2O3 | 4% (v/v), 25 nm | DI water, engine oil, glycerine, EG |
Numerical work:
|
| Abdollahi et al. [132] | 200–700 | Al2O3 | 1.0% (v/v), 2.5% (v/v), 5.0% (v/v) | DI water |
Numerical work:
|
| Vinoth et al. [131] | 300–850 | Al2O3 | 0.25% (v/v) | Distilled water |
Experimental work:
|
| Duangthongsuk et al. [130] | NA | SiO2 | 0.3% (v/v), 0.6% (v/v), 0.8% (v/v) | DI water |
Experimental work:
|
| Arani et al. [129] | 500–2,000 | SWCNT | 4% (v/v), 8% (v/v) | Distilled water |
Numerical work:
|
| Thansekhar and Anbumeenakshi [128] | 55–90 | Al2O3, SiO2 | 0.1% (v/v), 0.25% (v/v) | DI water |
Experimental work:
|
| Sarafraz et al. [127] | 100–1,400 | MWCNT | 0.05% (w/v), 0.1% (w/v), 0.15% (w/v) | DI water |
Experimental work:
|
| Snoussi et al. [126] | 100–1,000 | Al2O3 Cu | 1.0% (v/v), 2.0% (v/v) | DI water |
Numerical work:
|
| Abubakar et al. [124] | 140, 700, 1,400, | Fe3O4 | 0.4% (v/v), 0.6% (v/v), 0.8% (v/v) | Distilled water |
Numerical work:
|
| Sivakumar et al. [125] | 100–1,300 | CuO, Al2O3 | NA | EG, Distilled water |
Experimental and numerical work:
|
However, Razali et al. [138] found that the performance of Al2O3/water decreased when used for a longer period because of the changes in the crystallite size of Al2O3 that was affected by its growth during the heating process. To overcome the crystallisation problem upon heating, Sivakumar et al. [125] studied on another potential additive, CuO nanoparticle. The performance of Al2O3 nanoparticle was compared with CuO nanoparticle using different base fluids such as EG and pure water. The results showed that the heat transfer coefficient of CuO/EG was higher than Al2O3/water and CuO/water that could be attributed to the higher viscosity and density of EG as the base fluid. However, the findings contradicted with that reported by Alfaryjat et al. [133] where water was found to perform better as a base fluid compared to EG. Besides the type of base fluid, the volume concentration of nanoparticle also affects the heat transfer performance negatively or positively. The most common volume concentration used in the previous studies was less than 2%. Chabi et al. [139] claimed that nanofluid with a high volume concentration tends to reduce the heat transfer coefficient because of the rapid formation of sedimentation besides reducing the nanofluid stability. The statement by Chabi et al. [139] was further supported by Manay and Sahin [140] where a volume concentration of less than 2.0% of TiO2 nanoparticle increased the overall thermal efficiency with an increasing volume concentration of nanoparticle. However, at a higher volume concentration, the heat transfer performance decreased. Increasing the volume concentration also significantly affected the thermal resistance because of particle deposition. Furthermore, the reduction in thermal resistance was reported when nanoparticle of an average diameter of 25 nm was used.
All studies reviewed in this paper supported the fact that nanoparticles that dispersed in a base fluid can enhance the heat transfer performance in an MCHS. However, at a higher volume concentration of nanofluid, high-pressure drop penalty dominated the thermal performance enhancement. Majority of studies focused on the use of water as the base fluid and Al2O3 nanoparticle as the additive. According to Elsevier database from 2015 to 2017, 87% of the published papers used water (DI water or distilled water) as a base fluid for nanofluid preparation in MCHS application as shown in Figure 3. Besides that, higher application of Al2O3 nanoparticle as additives in an MCHS was noted compared to the use of MWCNT, TiO2, Ag, SiO2, SWCNT, Cu, CuO, Zn, and SiC as additives as illustrated in Figure 4. Nanofluid concentration was found to significantly affect the overall performance of an MCHS. The use of optimum nanofluid concentration helps to minimise the penalty of pressure drop in an MCHS. Figure 5 presents the concentration of nanofluid, which was commonly used in the MCHS application. Most of the studies used nanoparticle dispersed in a base fluid with maximum concentrations of 2% (v/v) for the MCHS application.
Type of base fluids commonly used in nanofluid preparation.
Type of nanoparticles commonly used in nanofluid preparation.
Distribution of nanoparticle concentration commonly used in MCHS applications.
3 Future prospects of MCHS with passive method and Moore’s Law
Rapid development of MCHS in the thermal management of electronic device is expected to be continued until the year of 2040. According to Moore’s law used by Intel for transistor count observation, the number of transistors has doubled every 24 months [141,142,143,144,145] as shown in Figure 6 [142]. There are three measures considered in the increasing number of transistors per microchips, such as shrinking the size of the single transistor (scaling), increasing the microchip area, and improving circuit and device design. As we can see here, scaling down the individual transistor and microchip size in improving the microprocessor performance (as shown in Figure 7) [146,147] has caused detrimental effects leading to high heat dissipation. This is one of the biggest challenges in the microchip development [148]. As a consequence, urgent needs for the advanced cooling device are rising by years.
![Figure 6
Evolution of transistor count according to Moore’s law [142].](/document/doi/10.1515/ntrev-2020-0094/asset/graphic/j_ntrev-2020-0094_fig_006.jpg)
Evolution of transistor count according to Moore’s law [142].
![Figure 7
Evolution of microchip size [146,147].](/document/doi/10.1515/ntrev-2020-0094/asset/graphic/j_ntrev-2020-0094_fig_007.jpg)
In recent years, the microchip development has saturated because the size of 3D transistor reached its physical limit because of the scaling down effect. In 2019, Li et al. [149] proposed a 2D transistor, which enabled to extend the Moore’s law, to replace the 3D transistor that can shrink further. Therefore, heat will dissipate exponentially in the future.
As mentioned in the previous section, a microchannel is a spin-off technology that can be implemented in many fields. Why microchannel becomes the first choice compared to nanochannel in many application? This is because microchannel can be fabricated by micromachining process with the high-quality output. In contrast, nanochannel is fabricated by laser machining process with some issues output such as cracking, ripple formation, shape deformation, and debris deposition [150,151].
MCHS has a bright future in many applications. Based on the market research report [152], MCHS is one of the cooling technologies that expected to grow at the highest compound annual growth rate (CAGR) in the global thermal management market for the global forecast of the year 2025.
4 Conclusion
Current designs of MCHSs with novel channel designs that could result in a high heat transfer rate with acceptable pressure drop penalty were comprehensively reviewed in this study. This review is expected to provide a better understanding of microchannel design development and the effects of geometry design on hydrothermal performances such as heat transfer and fluid flow characteristics. These two characteristics are critical in the design development of MCHS to achieve high heat transfer rate with a low cost concerning energy consumption. Besides that, the effect of widely used nanoparticle concentration, type, and base fluid selection on nanofluid performance and stability in the MCHS application have been discussed and summarised. Following are the main conclusions that can be derived from this review:
Hydraulic performance of hybrid passive method of MCHS was better than single passive method of MCHS over a wide range of Re. Rapid development of vortices flow formation at low Re increases fluid mixing in the hybrid MCHS with low pressure drop penalty, and thus reduces pumping power consumption.
Combination of more than one passive method in a hybrid MCHS can achieve a high heat transfer performance with minimal pressure drop penalty. However, sometimes the minimal pressure drop penalty can be very high because of the friction factor influenced by geometry structure. Therefore, the optimum hydrothermal performance should also consider the friction factor to prevent any leakage in the pipelines or channels.
The key strategy to improve the heat transfer performance is to increase the fluid mixing and redevelop the boundary layer in the laminar flow, which can be achieved by changing the flow profile. The out-of-flow mixing method is the best method that can provide a high degree of flow mixing; however, this method results in a high-pressure drop. Alternatively, the secondary flow generated by secondary channel geometry can be very useful to increase the degree of flow mixing without a high-pressure drop.
The surface roughness method is the best method to reduce friction factor and at the same time can enhance the heat transfer rate by providing a larger heat transfer area. However, this method is ineffective if applied as a single method, as the recirculation zone in the cavity area causes the fluid to remain in the cavity and does not remove the heat flux efficiently. By combining the method with flow disruption method, the fluid in the cavity area can be replaced easily by guiding the mainstream at the centre channel to the cavity area using the rib geometry.
The physical properties of base fluid such as thermal conductivity, dynamic viscosity, heat capacity, and density significantly affected the nanofluid performance. Pure water such as DI water and distilled water were widely used as the base fluid for nanofluid preparation for the MCHS application.
Each nanoparticle has a limited range of volume concentration utilisation. Optimum volume concentration application can increase the heat transfer performance with a minimal pressure drop in the MCHS application. However, higher nanoparticle volume concentration in a base fluid can increase the thermal resistance because of particle deposition and thus reducing the overall thermal performance of the MCHS. Nanoparticle solution with a volume concentration of less than 2% is suggested. Besides, Al2O3 nanoparticle is recommended to be used as an additive, which is commonly used in MCHS applications for nanofluid preparation.
Acknowledgments
The authors would like to thank Malaysia-Japan International Institute of Technology (MJIIT) and Universiti Teknologi Malaysia for supporting this research. This work was funded by the FRGS-MSRA grant [R.K130000.7843.5F273], LRGS grant [R.K130000.7843.4L891], and TAKASAGO grant [R.K130000.7343.4B472].
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Conflict of interest: The authors declare no conflicts of interest regarding the publication of this paper.
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