US20100132924A1

US20100132924A1 - Cooling device for electronic components

Info

Publication number: US20100132924A1
Application number: US12/597,774
Authority: US
Inventors: Kim Choon Ng; Christopher Robert Yap; Mark Aaron Chan
Original assignee: National University of Singapore
Current assignee: National University of Singapore
Priority date: 2007-04-27
Filing date: 2008-04-24
Publication date: 2010-06-03
Also published as: WO2008133594A2; WO2008133594A3

Abstract

A cooling device (10) including a chamber having an evaporator section (34) and a condenser section, the condenser section extending from a periphery of the evaporator section (34).

Description

FIELD OF THE INVENTION

The present invention relates to cooling devices for electronic components.

BACKGROUND OF THE INVENTION

Advances in photolithography have enabled dense packing of transistors on a chip. This has in turn resulted in an increase in the amount of heat generated per unit area of the chip, in some instances, exceeding the thermal design limit of conventional cooling devices. There is thus a need for a cooling device that is capable of dissipating the waste heat generated by electronic components effectively.

SUMMARY OF THE INVENTION

In a first aspect, the invention provides a cooling device comprising a chamber having an evaporator section and a condenser section, the condenser section extending from a periphery of the evaporator section.
A compact cooling device is thus achieved with the back-to-back configuration of the evaporator and condenser sections. The compact arrangement of the cooling device results in almost zero pressure differential between the evaporator and condenser sections (and hence, a uniform temperature distribution in the chamber) for fluid flow within the chamber. This reduces the overall thermal resistance of the cooling device.
A working fluid may be arranged to absorb and transfer heat energy. The working fluid may have a latent heat of vaporisation of greater than or equal to about 1500 joules per gram (J/g). The working fluid may include a plurality of nanoparticles, the nanoparticles having a thermal conductivity of between about 200 Watt per metre Kelvin (W/m·K) to about 400 W/m·K.
The condenser section may comprise a plurality of condenser fins extending from the periphery of the evaporator section.
A plurality of convective fins may extend from surfaces of the condenser fins. This facilitates heat transfer from the condenser fins.
The convective fins may extend between adjacent ones of the condenser fins. This enhances the compactness of the cooling device.
A fan may be arranged to direct a flow of air at the convective fins. This enhances convective heat transfer from the convective fins and thus improves the heat rejection rate.
An air guide may be arranged to direct the flow of air from the fan to the convective fins. This enhances convective heat transfer from the convective fins.
The convective fins may be integrally moulded to the surfaces of the condenser fins. This reduces thermal contact resistance and thus increases thermal conductivity between the convective fins and the condenser fins.
A capillary element may be arranged to facilitate return of condensate from the condenser section to the evaporator section. This facilitates return of the condensate to the evaporator section regardless of the orientation of the cooling device.
A base portion of the condenser section may be at an angle to a base portion of the evaporator section. This facilitates return of the condensate to the evaporator section by gravitational means.
The evaporator section may have a capillary structure. The capillary structure enhances boiling in the evaporator section as it creates a micro-flow situation within the evaporator section by drawing heated working fluid away from the base portion of the evaporator section and fresh working fluid in the form of condensate from the condenser fins to the evaporator section. The degree of superheat (i.e. the difference between the temperature of the base portion of the evaporator section and the saturation temperature of the working fluid) is thus reduced. The capillary structure may comprise an open cell foam metal. The open cell foam metal may be bonded to a base portion of the evaporator section by diffused bonding or brazing to reduce thermal contact resistance and thus improve thermal conductivity between the base portion of the evaporator section and the capillary structure. The capillary structure may have a pore density of from about 8 to about 52 pores per centimetre (ppc) and more particularly, a pore density of about 16 ppc.
Other aspects and advantages of the invention will become apparent from the following detailed description, taken in conjunction with the accompanying drawings, illustrating by way of example the principles of the invention.

BRIEF DESCRIPTION OF THE DRAWINGS

Embodiments of the invention will now be described, by way of example only, with reference to the accompanying drawings, in which:

FIG. 1A is an exploded top perspective view of a cooling device in accordance with one embodiment of the present invention;

FIG. 1B is an exploded bottom perspective view of the cooling device of FIG. 1A;

FIG. 2A is an assembled top perspective view of the cooling device of FIG. 1A;

FIG. 2B is an assembled bottom perspective view of the cooling device of FIG. 1B;

FIG. 3 is a schematic cross-sectional view of the cooling device taken along section lines A-B and B-C in FIG. 2B; and

FIG. 4 is a graph showing the transient performance of a cooling device with convective fin thickness of 0.25 millimetres (mm).

DETAILED DESCRIPTION OF EXEMPLARY EMBODIMENTS

The detailed description set forth below in connection with the appended drawings is intended as a description of the presently preferred embodiments of the invention, and is not intended to represent the only form in which the present invention may be practiced. It is to be understood that the same or equivalent functions may be accomplished by different embodiments that are intended to be encompassed within the scope of the invention.
A cooling device 10 for an electronic component (not shown) such as, for example, a microprocessor or central processing unit (CPU) of a computer will now be described below with reference to FIGS. 1A, 1B, 2A, 2B and 3.
The cooling device 10 comprises a fan 12 mounted to an air guide 14, a hermetically sealed chamber 16 having a plurality of convective fins 18 and a sleeve 20, and an attachment member 22 in the form of a leg support. The hermetically sealed chamber 16 is provided with a working fluid 24 to absorb and transfer heat energy.
The fan 12 is arranged to direct a flow of air at the convective fins 18 and the air guide 14 is arranged to direct the flow of air from the fan 12 to the convective fins 18. In the present embodiment, the fan 12 is mounted to the air guide 14 by a plurality of beams 26 extending from the stator 28 of the fan 12 to the air guide 14, the fan 12 being positioned such that a separation is maintained between the rotor 30 of the fan 12 and the hermetically sealed chamber 16. As can be seen from FIGS. 2B and 3, the air guide 14 is spaced from the outermost convective fin 18 to allow airflow through.
As shown in FIG. 3, the hermetically sealed chamber 16 is mounted to the air guide 14 via a clipping mechanism 32. The hermetically sealed chamber 16 comprises an evaporator section 34 and a condenser section comprising a plurality of condenser fins 36 extending from a periphery of the evaporator section 34. In the present embodiment, a base portion of the condenser section is slanted at an angle θ to a base portion of the evaporator section 34. This facilitates return of condensate 38 from the condenser fins 36 to the evaporator section 34 by gravitational means. The base portion of the evaporator section 34 in contact with the electronic component may be polished to reduce thermal contact resistance and thus improve heat transmission from the electronic component to the evaporator section 34.
In the present embodiment, the evaporator section 34 is provided with a capillary structure 40. The capillary structure 40 may, for example, be provided in the form of an open cell foam metal made of copper; aluminium or other highly thermally conductive material and may be bonded to the base portion of the evaporator section 34 by, for example, diffused bonding or brazing to reduce thermal contact resistance and thus improve thermal conductivity between the base portion of the evaporator section 34 and the capillary structure 40. The capillary structure 40 may have a pore density of from about 8 to about 52 pores per centimetre (ppc) (equivalent to from about 20 to about 130 pores per inch (ppi)) and more particularly, a pore density of about 16 ppc.
The condenser fins 36 are hollow and of a certain thickness with one end open towards the evaporator section 34 to allow working fluid 24 in the vapour phase to enter from the evaporator section 34 and condense even on its outermost end, thereby spreading heat effectively and achieving a more uniform temperature distribution. In the present embodiment, each of the condenser fins 36 is provided with a capillary element 42 such as, for example, a cotton thread or braided copper wires to facilitate return of the condensate 38 from the condenser section to the evaporator section 34 at various orientations of the cooling device 10. A small groove (not shown) may be provided on the slanted bottom of the condenser fin 36 for receiving the capillary element 42 in the form of a capillary pumping thread.
The working fluid 24 in the hermetically sealed chamber 16 may be a single or multi component condensable fluid. Examples of the working fluid 24 include, but are not limited to, water, dielectric fluids such as, for example, hydrofluoroethane (HFE) or fluorocarbon (FC) in the refrigerant series, or any other fluid or combination of fluids with a high latent heat of evaporation (more particularly, a latent heat of vaporisation of greater than or equal to about 1500 joules per gram (J/g)) and which can permeate the foam metal material in the evaporator section 34. The working fluid 24 may be provided with a plurality of nanoparticles to augment the effective thermal conductivity of the working fluid 24. The nanoparticles may have a thermal conductivity of between about 200 Watt per metre Kelvin (W/m·K) to about 400 W/m·K. An optimum amount of working fluid 24 is provided in the hermetically sealed chamber 16 such that the capillary structure 40 is always in contact with and saturated with the working fluid 24 to prevent drying out and to ensure wetting of the capillary structure 40 throughout the operation of the cooling device 10. The optimum amount of working fluid 24 in the hermetically sealed chamber 16 is determined based on the amount of heat generated by the electronic component. A larger quantity of working fluid 24 is required in instances where more heat is generated to prevent dry out. The working fluid 24 in the hermetically sealed chamber 16 may be in its liquid-vapour saturated state to facilitate conversion of the working fluid 24 into its vapour form as this enhances heat dissipation by the cooling device 10. To achieve the liquid-vapour saturated state, the hermetically sealed chamber 16 may be in a pressurized or vacuum state depending on the working fluid 24 employed. In other words, the pressure in the hermetically sealed chamber 16 is dependent on the saturation properties of the working fluid 24. As an example, the hermetically sealed chamber 16 may be in a vacuum state when water is employed as the working fluid 24.
As can be seen from FIGS. 1A, 1B, 2A and 2B, the convective fins 18 extend from surfaces of the condenser fins 36 and between adjacent ones of the condenser fins 36. The convective fins 18 may be integrally moulded to the surfaces of the condenser fins 36 to reduce thermal contact resistance and thus increase thermal conductivity between the convective fins 18 and the condenser fins 36. In such an embodiment, the hermetically sealed chamber 16 and the convective fins 18 may be made from a single piece of material such as, for example, copper, aluminium or other highly thermally conductive material that is compatible with the working fluid 24, via a moulding or machining process.
As shown in FIG. 1B, the sleeve 20 is provided with threaded holes 44 to receive screws (not shown) for securing the attachment member 22 to the evaporator section 34.
The attachment member 22 provides an interface between the hermetically sealed chamber 16 and the electronic component. As can be seen from FIGS. 1B and 2B, the attachment member 22 is provided with first threaded holes 46 corresponding to the threaded holes 44 on the sleeve 20 and second threaded holes 48 for attachment, for example, to a motherboard of a computer. The attachment member 22 is secured to the sleeve 20 by passing screws (not shown) through respective pairs of the threaded holes 44 on the sleeve 20 and the first threaded holes 46 on the attachment member 22, and fastening the screws. The attachment member 22 is similarly secured, for example, to the motherboard of the computer. The leg support 22 may be of any configuration depending on the layout of the surface to which the cooling device 10 is to be secured.
The operation of the cooling device 10 will now be described with reference to FIG. 3 which shows a schematic cross-sectional view of the cooling device 10 taken along section lines A-B and B-C in FIG. 2B. Accordingly, the left side of FIG. 3 shows the air flow through the convective fins 18 of the cooling device 10 and the right side illustrates circulation of the working fluid 24 within the hermetically sealed chamber 16.
When in use, an electronic component generates and dissipates heat. The heat from the electronic component is taken in at the base portion of the evaporator section 34 and raises the temperature of the base portion of the evaporator section 34.
Boiling of the working fluid 24 inside the evaporator section 34 of the hermetically sealed chamber 16 occurs as heat is taken from the base portion of the evaporator section 34 by the working fluid 24 as latent heat of vaporization. The capillary structure 40 in the evaporator section 34 enhances boiling of the working fluid 24 as it creates a micro-flow situation within the evaporator section 34 by drawing the heated working fluid 24 away from the base portion of the evaporator section 34 and fresh working fluid 24 in the form of condensate 38 from the condenser fins 36 to the evaporator section 34 (see arrow 50 in FIG. 3). The degree of superheat (i.e. the difference between the temperature of the base portion of the evaporator section 34 and the saturation temperature of the working fluid 24) is thus reduced.
Evaporation of the working fluid 24 in the capillary structure 40 and the escape of vapour bubbles from the liquid bulk in the evaporator section 34 to the liquid-vapour interface result in the creation of a higher pressure region in the evaporator section 34, forcing the vapour in the evaporator section 34 to a relatively lower pressure region in the hollow condenser fins 36 (see arrows 52 in FIG. 3).
Condensation occurs on inner walls 54 of the condenser fins 36 as the inner walls 54 are at a lower temperature than the vaporised working fluid 24 in the condenser fins 36. Accordingly, the working fluid 24 is collected as condensate 38 on the inner walls 54 of the condenser fins 36. Depending on the orientation of the cooling device 10, the condensate 38 may be returned to the evaporator section 34 by gravitational means where the condenser fins 36 are at a higher elevation than the evaporator section 34, by capillary action via the capillary element 42 (particularly when the condenser fins 36 are at a lower elevation than the evaporator section 34) or a combination of both.
Heat in the form of latent heat of condensation transmitted from the working fluid 24 to the condenser fins 36 is conducted to the convective fins 18 which provide a large surface area for heat dissipation. Heat is convected away from the convective fins 18 to ambient space by the flow of air from the fan 12 (see arrows 56 in FIG. 3). More particularly, the fan 12 mounted at the top of the main body of the cooling device 10 and located inside the opening end of the air guide 14 forces air to flow through the array of convective fins 18. The air guide 14 constricts and guides the air flow into the array of convective fins 18.
Although illustrated in an upright position in FIGS. 1A, 1B, 2A, 2B and 3, it should be understood that the cooling device of the present invention is not limited to a particular orientation. Rather, the cooling device may be used in any orientation as long as the capillary structure is in contact with the bulk working fluid in liquid phase for capillary pumping.

Simulation Results

A numerical simulation was carried out to gauge the performance of the cooling device 10 using water as the working fluid 24 and copper for the chamber 16 and the convective fins 18. Convective cooling is provided by an axial flow fan 12 with a diameter of 92 millimetres (mm) and which generates an air flow rate of about 11.80 litre per second (l/s) (equivalent to 25 cubic feet per minute (CFPM)). The governing equations for the simulation (i.e. conservation equations for energy and mass provided below) are solved simultaneously using the International Mathematical and Statistical Library (IMSL) running on a Fortran PowerStation platform where the convergence criterion is a tolerance of 10⁻⁹. The steady state performance parameters of the cooling device 10 subjected to 10 watts per square centimetre (W/cm²) of heat flux from a CPU measuring 3 by 3 square centimetre (cm²) at ambient air temperature of 30 degrees Celsius (° C.) are shown in Table 1 below.
TABLE 1

Thickness No. of CPU Air Flow

of Conv. Conv. Temperature R_conv R_total Rate P_drop

Fins (mm) Fins (° C.) (K/W) (K/W) (l/s) (Pa)

0.15 18 49.1 0.123 0.218 11.89 24.99

0.25 17 48.1 0.116 0.206 11.94 24.91

0.50 14 48.7 0.123 0.213 10.48 26.71

0.75 12 49.9 0.136 0.226 9.34 28.18

1.00 10 51.2 0.151 0.241 9.25 28.29

wherein R_convrepresents the convective thermal resistance between the convective fins 18 and ambient air measured in kelvin per watt (K/W), R_totalrepresents the total thermal resistance of the cooling device 10 and P_droprepresents the pressure drop across the convective fins 18 and is measured in pascal (Pa).
The optimal performance for convective fins 18 of various thicknesses is shown in Table 1. As can be seen from Table 1, the cooling device 10 with the thinnest convective fins 18 measuring 0.25 mm outperforms all others by maintaining the CPU temperature at a mere 48.1° C. with an overall but optimal thermal resistance of 0.206 K/W as it allows placement of an optimum number of convective fins 18 while allowing an adequate amount air flow from the axial flow fan 12 and thus achieves the highest convective heat transfer. However, as convective fin thickness is decreased further, it was observed that the benefits of increasing convective surface area by reducing fin thickness is somewhat offset by a drop in the fin efficiency for the thinner fins. It is concluded therefore that fin strength is also a design consideration.
Referring now to FIG. 4, the transient performance of the cooling device 10 with convective fin thickness of 0.25 mm is shown. The data for the graph of FIG. 4 was obtained by performing a simulation test on a cooling device 10 with seventeen (17) convective fins 18, each convective fin having a thickness of 0.25 mm, a height of 30 mm and a length of 40 mm, the cooling device 10 being subjected to varying heat flux loads ranging from 10 W/cm²to 25 W/cm²from a 9 cm²CPU footprint. Temperatures of the CPU, the base portion of the evaporator section 34, the saturated working fluid 24 in the hermetically sealed chamber 16, and the inner walls 54 of the condenser fins 36 (in particular, at the base portion of the condenser fin array) were measured and plotted against time till a steady state was reached.
In all the simulated cases, the CPU chip temperature did not exceed 70° C.—the thermal design limit of the CPU—even at a CPU heat flux of 25 W/cm²(i.e. a total of 225 watts (W)).
The following were observed from the results of the simulation:
(i) The thermal resistance between the convective fins 18 and ambient air is highest for all simulated heat flux loads.
(ii) The next highest thermal resistance in the cooling device 10 is the superheat for boiling. In this regard, it was observed that the heat transfer coefficient of boiling increases with increasing surface heat flux at the base portion of the evaporator section 34. Thus, the corresponding increase in boiling superheat is minimal.
(iii) The temperature differential between the CPU and the base portion of the evaporator section 34 was observed to increase from a small differential of 2.3° C. to a high of 5.8° C. This indicates that both the heat spreading and interface resistances increase with increasing heat flux loads.
(iv) The temperature differential between the condenser wall and the saturated working fluid 24 was observed to increase from a small differential of 0.5° C. to 1.5° C. This small increase is acceptable and indicates that condensation is an efficient process for heat rejection.
Observations (i) and (ii) are consistent with existing case studies.
The heat spreading problem may be alleviated by using an evaporator with a smaller base area. In regard to the boiling resistance, the thermal conductivity of the working fluid 24 may be enhanced by adding nanoparticles to the working fluid 24 to reduce the boiling superheat. The size of the convective fins 18 may be varied according to the space available to accommodate future higher heat fluxes.

Equations for Simulation

Energy Balance at Base Portion of Evaporator Section

The energy balance at the base portion of the evaporator section 34 made of copper, taking into consideration the heat given out by the CPU chip and the heat extracted by evaporation of the working fluid 24, may be modelled with the following equation:
$\begin{matrix} [{(M c_{p})}^{base}] \frac{\partial T^{base}}{\partial t} = - q_{flux}^{evap} A^{evap} + q_{flux}^{chip} A^{chip} & (1) \end{matrix}$
wherein (Mc_p)^baserepresents the thermal capacity of the base portion of the evaporator section 34,
$\frac{\partial T^{base}}{\partial t}$
represents the rate of change of the temperature at the base portion of the evaporator section 34 over time, q_flux ^chiprepresents the heat flux input from the heat source, A^chiprepresents the footprint area of the heat source, A^evaprepresents the bottom surface area of the evaporator section 34, and q_flux ^evaprepresents the heat taken in by the working fluid 24 for evaporation.

Energy Balance of Working Fluid in Hermetically Sealed Chamber

The hermetically sealed chamber 16 includes a porous structure 40 and a saturated working fluid 24. Latent heat is taken in by the working fluid 24 from the superheated surface of the base portion of the evaporator section 34, and is transported and rejected to the finned condenser section. The energy balance of the working fluid 24 may be modelled with the following equation:
$\begin{matrix} [{(M c_{p})}^{eff}] \frac{\partial T^{chamber}}{\partial t} = - h_{g} (T^{evap}) \frac{\partial m_{vap}}{\partial t} + h_{f} (T^{cond}) \frac{\partial m_{liq}}{\partial t} & (2) \end{matrix}$
wherein (Mc_p)^effrepresents the effective thermal capacity of the working fluid 24 and is obtained by summing the thermal capacity of the working fluid 24 in the vapour, liquid and phases (i.e. (Mc_p)^eff=(MCp)^vapor+(MCp)^liquid+(MCp)^foam),
$\frac{\partial T^{chamber}}{\partial t}$
represents the rate of change of the temperature of the vapour space in the chamber 16 over time, h_gand h_frepresent the respective enthalpies of the working fluid 24 in vapor and liquid phases at a given temperature, T_evaprepresents the temperature of the section 34, T^condpresents the temperature of the condenser section,
$\frac{\partial m_{vap}}{\partial t}$
represents the mass flow rate of the working fluid 24 in vapour phase and
$\frac{\partial m_{liq}}{\partial t}$
represents the mass flow rate of the condensate 38.
The Rohsenow Correlation, adjusted for sub-atmospheric conditions as shown in equations (3) and (4) below, is used to predict the heat flux taken in by the working fluid 24 for boiling at a certain superheat at sub-atmospheric conditions in the presence of the porous structure 40:
$\begin{matrix} T^{surface} = T^{evap} + \overset{\overset{Rohsenow Correlations}{}}{{(\frac{C_{sf} h_{fg} \Pr_{l}^{s}}{c_{pf}}) [\frac{q_{flux}^{evap}}{μ_{l} h_{fg}} \sqrt{\frac{σ}{g (ρ_{l} - ρ_{g})}}]}^{0.33}} {(\frac{P}{P_{at m}})}^{m} {(\frac{A_{wetted}}{A_{base}})}^{n} & (3) \\ q_{flux}^{evap} = {(\frac{μ_{l} h_{fg}}{\sqrt{\frac{σ}{g (ρ_{l} - ρ_{g})}}}) [\begin{matrix} (T^{surface} - T^{evap}) (\frac{c_{pf}}{C_{sf} h_{fg} \Pr_{l}^{s}}) \\ {(\frac{P_{at m}}{P})}^{m} {(\frac{A_{base}}{A_{wetted}})}^{n} \end{matrix}]}^{(1 / 0.33)} & (4) \end{matrix}$
wherein s is 1, m is 0.293, n is −0.0984, T^surfacerepresents the temperature of the base portion of the evaporator section 34, C_Sf, an empirical constant related to the fluid-heater surface combination, is 0.0132, h_fgrepresents the latent heat of evaporation of the working fluid 24, Pr_lrepresents the Prantl Number, C_pfrepresents the heat capacity of the working fluid 24 in liquid phase, μ_lrepresents the viscosity of the working fluid 24 in liquid phase, σ represents surface tension, g represents acceleration due to gravity, ρ_lrepresents the density of the working fluid 24 in liquid phase, ρ_grepresents the density of the working fluid 24 in vapour phase, P represents the pressure in the evaporator section 34, P_atmrepresents atmospheric pressure, A_wettedrepresents the wetted surface area of the porous structure 40 and A_baserepresents the surface area of the base portion of the evaporator section 34.
The Thin Film Condensation Correlation is used to predict the heat transfer coefficient h_iof the condenser section as shown in equation (5) below on the assumption that the walls of the condenser section are acting as individual plates:
$\begin{matrix} h_{i} = 0.934 [\frac{ρ_{l} (ρ_{l} - ρ_{g}) {gh}_{fg} k^{3}}{μ (T_{v} - T_{cond}) L_{fin}}] & (5) \end{matrix}$
wherein k represents the thermal conductivity of the working fluid 24 in liquid phase, μ represents the viscosity of the working fluid 24 in liquid phase, T_vrepresents the temperature of the working fluid 24 in vapour phase in the hermetically sealed chamber 16 and L_finrepresents the length of the condensate 38 flow path.

Mass Balance

Mass balance was performed on the hermetically sealed chamber 16 to account for all the liquid and vapour mass inside the chamber 16. This provides information for determining the optimum amount of working fluid 24 to fill the hermetically sealed chamber 16 with to prevent drying out and to ensure wetting of the capillary structure 40 throughout the operation of the cooling device 10. The mass balance was modelled with the following equations:
$\begin{matrix} \frac{\partial M^{vapor}}{\partial t} = {\dot{m}}_{evaporation} - {\dot{m}}_{condensation} & (6) \\ \frac{\partial M^{liquid}}{\partial t} = {\dot{m}}_{condensation} - {\dot{m}}_{evaporation} & (7) \end{matrix}$
wherein
$\frac{\partial M^{vapor}}{\partial t}$
represents the net flow rate of the vapour in the chamber 16,
$\frac{\partial M^{liquid}}{\partial t}$
represents the net flow rate of the liquid in the chamber 16, in {dot over (m)}_evaporationrepresents the rate of evaporation and {dot over (m)}_condensationrepresents the rate of condensation.
The rate of evaporation {dot over (m)}_condensationand the rate of condensation {dot over (m)}_condensationmay be determined with the following equations:
$\begin{matrix} {\dot{m}}_{evaporation} = \frac{Q_{flux}^{evap}}{h_{fg} (T_{chamber})} & (8) \\ {\dot{m}}_{condensation} = \frac{h_{i} (T^{chamber} - T^{cond}) A_{cond}}{h_{fg} (T^{chamber})} & (9) \end{matrix}$
wherein Q_flux ^evaprepresents the heat flux at the base portion of the evaporator section 34 and is equivalent to q_flux ^evapdefined above, T_chamberrepresents the temperature in the vapour space of the hermetically sealed chamber 16 and A_condrepresents the surface area of the inner walls 54 of the condenser fins 36.

Energy Balance at Condenser Section

An energy balance was performed on the condenser fins 36. Latent heat is transferred to the inner walls 54 of the condenser fins 36 when the working fluid 24 condenses on the inner walls 54. The heat transmitted to the condenser fins 36 is then conducted to the convective fins 18 where it is transferred to ambient air by forced convection. The energy balance at the condenser section may be modelled with the following equation:
$\begin{matrix} [{(M c_{p})}^{cond}] \frac{\partial T^{cond}}{\partial t} = {\dot{m}}_{condensation} h_{fg} (T^{chamber}) - \frac{(T_{cond} - T_{air})}{R_{hs}} & (10) \end{matrix}$
wherein (Mc_p)^condrepresents the heat capacity of the condenser section,
$\frac{\partial T^{cond}}{\partial t}$
represents the rate of change of the temperature of the inner walls 54 of the condenser fins 36 over time, T_airrepresents ambient temperature, and R_hsrepresents the thermal resistance between the convective fins 18 and ambient air.
As is evident from the foregoing discussion, the present invention provides a compact cooling device with reduced heat spreading resistance for intense heat emitting surfaces such as, for example, a microprocessor or CPU of a desktop computer, as well as surfaces with high localised heat fluxes. The cooling device includes: an evaporator section where working fluid evaporates at its saturation pressure, extracting its latent heat of vaporization from the intense heat surface; a condenser section to reject latent heat from incoming working fluid in vapour phase from the evaporator section; a capillary structure located in the evaporator section to circulate the working fluid in liquid phase from the condenser section back to the evaporator section; an array of convective fins as a means of extending the heat sink surface area; and a fan to blow air through the array of convective fins, thereby enhancing heat rejection to ambient air. Advantageously, the evaporator section, the condenser section, the convective fins, and the fan can all be integrated into one compact device. By using a two-phase system, the overall size of the cooling device and heat spreading resistance can be significantly reduced, increasing the overall heat transfer capacity of cooling device.
While the preferred embodiments of the invention have been illustrated and described, it will be clear that the invention is not limited to these embodiments only. Numerous modifications, changes, variations, substitutions and equivalents will be apparent to those skilled in the art without departing from the scope of the invention as described in the claims.
Further, unless the context dearly requires otherwise, throughout the description and the claims, the words “comprise”, “comprising” and the like are to be construed in an inclusive as opposed to an exclusive or exhaustive sense; that is to say, in the sense of “including, but not limited to”.

Claims

1. A cooling device, comprising: a chamber having an evaporator section and a condenser section, the condenser section extending from a periphery of the evaporator section.

2. The cooling device according to claim 1, wherein the condenser section comprises a plurality of condenser fins extending from the periphery of the evaporator section.

3. The cooling device according to claim 2, further comprising a plurality of convective fins extending from surfaces of the condenser fins.

4. The cooling device according to claim 3, wherein the convective fins extend between adjacent ones of the condenser fins.

5. The cooling device according to claim 4, further comprising a fan arranged to direct a flow of air at the convective fins.

6. The cooling device according to claim 5, further comprising an air guide arranged to direct the flow of air from the fan to the convective fins.

7. The cooling device according to claim 3, wherein the convective fins are integrally moulded to the surfaces of the condenser fins.

8. The cooling device according to claim 1, further comprising a capillary element arranged to facilitate return of condensate from the condenser section to the evaporator section.

9. The cooling device according to claim 1, wherein a base portion of the condenser section is at an angle to a base portion of the evaporator section.

10. The cooling device according to claim 1, wherein the evaporator section has a capillary structure.

11. The cooling device according to claim 10, wherein the capillary structure comprises an open cell foam metal.

12. The cooling device according to claim 11, wherein the open cell foam metal is bonded to a base portion of the evaporator section by diffused bonding or brazing.

13. The cooling device according to claim 10, wherein the capillary structure has a pore density of from about 8 to about 52 pores per centimetre (ppc).

14. The cooling device according to claim 13, wherein the pore density of the capillary structure is about 16 ppc.

15. The cooling device according to claim 1, further comprising a working fluid arranged to absorb and transfer heat energy.

16. The cooling device according to claim 15, wherein the working fluid has a latent heat of vaporisation of greater than or equal to about 1500 joules per gram (J/g).

17. The cooling device according to claim 15, further comprising a plurality of nanoparticles in working fluid, the nanoparticles having a thermal conductivity of between about 200 Watt per metre Kelvin (W/m·K) to about 400 W/m·K.