Fig 1.
Geometrical and numerical configurations.
Top: View of one of the used meshes. The same locations and point of view are kept for both channels. Left: Straight channel (Height 1mm). Right Corrugated channel (Amplitude A0 = 90%). Vertical scale is magnified 100 times. Center: Top view schematic of the half-channel model. The vertical plane passing through the central axis of the channel is a plane of symmetry, so only half of the channel was simulated (W/2 = 25mm). Bottom: Side view schematic of the DDMC. The lower wall was fixed and subjected to a constant and uniform heat flux over the imprint of heating zone length (length, L: 50 mm, imposed heat power Γ: 125 W) facing the actuated upper wall. A negative pressure difference (ΔPs = Pout − Pin < 0) is imposed between outlet and inlet sections of the channel, deformed membrane drawn for 2 amplitudes (black and orange).
Table 1.
Mesh convergence for the straight channel.
Table 2.
Mesh convergence for the dynamic corrugated channel.
Fig 2.
Impact of thermo-dependent fluid properties and comparison with analytical results.
a) Global heat transfer coefficient (< h >) obtained using constant and variable thermo-physical properties as a function of the mass flow rate. (b) Evolution and validation of Nusselt number along x* for different Reynolds numbers in a rectangular channel heated by a constant heat flux on one side. Analytical results from Lee et al. [17] are obtained for an aspect ratio α ≤ 0.1.
Fig 3.
Instantaneous fields: (a) velocity field in a given corrugation, (b) velocity and (c) temperature fields in the symmetry plane. ΔPs = −50Pa, fr = 50Hz and A = 90%.
Fig 4.
(a) Time averaged pressure profile along the main flow axis (b) Instantaneous profiles at various points of time. The dips on the profile correspond to the constriction zone. Imposed pressure difference ΔPs = 0Pa, fr = 50Hz and A = 90%.
Table 3.
Performances of the dynamic corrugated heat exchanger for an amplitude of 89%.
Fig 5.
(a) Mass flow rate as a function of relative amplitude when ΔPs = −50Pa, (b) Variation in mass flow rate with imposed pressure difference for A = 95%, fr = 50Hz.
Fig 6.
(a) Global heat transfer coefficient (< h >), and (b) Mass flow rate as a function of relative amplitude for different ΔPs and fr = 50Hz.
Fig 7.
Global heat transfer coefficient.
< h > as a function of mass flow rate for (a) different relative amplitudes and ΔPs = −50Pa, and for (b) different ΔPs for a relative amplitude of 95% for the dynamic corrugated channel. fr = 50Hz. Comparison with flat and static channel is also presented.
Fig 8.
Performances of the dynamic corrugated heat exchanger.
Pumping power, pressure work and viscous loss for corrugated moving channel as a function of the imposed negative pressure differences for relative amplitude of A = 89% and fr = 50Hz.
Fig 9.
Performances of the dynamic corrugated heat exchanger.
(a) Mass flow rate and (b) Heat transfer coefficient for corrugated moving, static with and without lateral damping (log-log scales). Different imposed negative pressure differences for relative amplitude of A = 89% and fr = 50Hz.
Table 4.
Performances of the static corrugated heat exchanger with an amplitude of 89%.
Fig 10.
Performances of the dynamic corrugated heat exchanger.
Merit factor for corrugated moving, static with and without lateral damping (log-log scales). Different imposed negative pressure differences for relative amplitude of A = 89% and fr = 50Hz.